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BEARINGS 

AND 

THEIR   LUBRICATION 


Bearings 


and 


Their   Lubrication 


By  L.  P.  Alford,  M.  E. 

EDITOR,    AMERICAN    MACHINIST; 

MEMBER,    AMERICAN    SOCIETY   OF   MECHANICAL   ENGINEERS; 

MEMBER,    AMERICAN   SOCIETY    FOR.  TESTING   MATERIALS. 


ALLEGE 


U.  OFC. 

LIBRARY 


19  1 1 

Published  by  the 

American    Machinist 

McGraw-Hill  Book  Company,  Sole  Selling  Agents 

239  West  Thirty-ninth  Street,  New  York 

London,  E.  C,  6  Bouverie  Street  Berlin,  N.  W.  7.  Unter  der  Linden  71 


Copyright,  191  i,  by  the  American  Machinist 


7^t-^V£.wv- 


American  Machinist,  New  York,  U.S.  A. 


PREFACE 

The  aim  of  this  book  has  been  two  fold;  to  present  the  underlying  principles 
involved  in  the  design  of  all  classes  of  machinery  bearings  and  to  show  modern 
practice  in  the  construction  and  application  of  important  commercial  types. 
It  is  believed  to  be  the  first  treatment  of  these  subjects  ever  put  into  book  form. 

The  tendency  of  modern  machine  design  in  many  fields  is  toward  the  use  of 
high  speed  shafts  and  spindles  and  friction  reducing  forms  of  bearings.  There- 
fore the  present  seems  a  peculiarly  fitting  time  to  present  these  data  vdth  the 
purpose  of  making  them  permanently  useful  to  engineers,  designers,  draftsmen 
and  machinists;  in  fact  to  any  one  who  is  interested  in  any  way  with  machinery 
bearings  and  their  proper  lubrication  and  care. 

Many  of  the  data  have  never  before  been  published  in  any  form.  Many  of 
the  experiments  and  researches  quoted  are  from  European  sources  and  official 
transactions  of  technical  societies  and  are  not  generally  known  to  American 
readers. 

It  is  a  pleasure  to  acknowledge  the  many  courtesies  received  from  engineers 
and  machinery  manufacturers  and  the  valuable  information  furnished  by  them. 
Without  this  hearty  cooperation  it  would  have  been  impossible  to  have  produced 
this  work.  L.  P.  Alford. 

New  York  City. 

August,  191 1. 


CONTENTS 

Classification  of  Machinery  Bearings      


PART  I 

BEARINGS  WITH  SLIDING  CONTACT 

Section. 

I.  Sliding  Friction 5 

II.  Coefficients  OF  Friction  OF  Journal,  Collar,  Step  AND  Guide  Bearings  .    .  31 

III.  Materials  for  Bearings       41 

IV.  Allowable  Pressures,  Speeds  and  Temperatures      65 

V.  Design  of  Journal  Bearings      86 

VI.  Lubricants m 

VII.  Design  of  Flat  Sliding  Surfaces  and  Special  Bearings 132 

VIII.  Three  Important  Bearing  Inventions 139 

IX.  Typical  Designs  and  Constructions 141 

X.  Hints  on  the  Care  of  Bearings 166 

PART  II 

BEARINGS  WITH  ROLLING  CONTACT 

I.  Rolling  Friction  and  Factors  of  Design     173 

II.  Construction  of  Ball  Bearings 180 

III.  Typical  Designs  and  Mountings  for  Ball  Bearings 195 

IV.  Lubrication  of  Ball  Bearings       204 

V.  Roller  Bearings  with  Flexible  Rollers      209 

VI.  Radial  Roller  Bearings  with  Solid  Rollers 217 

VII.  Roller  Thrust  Bearings 221 


vu 


BEARINGS  AND  THEIR  LUBRICATION 


CLASSIFICATION  OF  MACHINERY  BEARINGS 

Machinery  bearings  are  the  parts  of  the  bed,  frame  or  other  members  that 
constrain  rotating  parts,  such  as  shafts  and  spindles.  They  are  divided  into  two 
general  classes,  journal  and  thrust,  depending  upon  the  direction  in  which  the 
load  acts.  In  the  journal  bearing  the  load  acts  at  right  angles  to  the  axis;  in 
thrust  bearings,  parallel  to  the  axis. 

The  bearing  surface  of  all  journal  bearings  is  necessarily  circular  in  cross- 
section.  In  profile  it  is  ordinarily  cylindrical,  but  may  be  conical,  spherical  or 
even  of  a  more  complicated  shape.  The  bearing  surface  of  thrust  bearings  is 
ordinarily  flat,  but  may  be  spherical,  conical,  or  shaped  to  the  curve  of  the 
tractrix. 

Bearings  are  also  divided  into  two  classes  by  the  kind  of  contact  between  the 
surfaces;  that  is,  bearings  having  sliding  contact,  or  ordinary  bearings,  and  bear- 
ings having  rolling  contact  or  ball  and  roller  bearings. 

The  bearing  surfaces  of  ball  bearings  both  radial  (journal)  and  thrust  are 
ordinarily  curved  races  with  a  circular  cross-section.  Less  commonly  the  sur- 
faces are  flat. 

The  bearing  surfaces  of  journal  roller  and  thrust  bearings  are  circular  in 
cross-section  and  either  cylindrical  or  conical  in  profile. 

The  accompanying  table  gives  a  classification  of  machinery  bearings. 
Flat  constraining  surfaces  for  sliding  machine  parts  having  rectilinear  move- 
ments have  no  common  name,  but  are  referred  to  as  ways,  guides,  and  the  like. 

The  relative  motion  of  all  these  constraining  members  and  the  members 
constrained  is  resisted  at  their  surfaces  of  contact  by  a  force  which  is  called  the 
force  of  friction.  Thus,  all  kinds  of  bearings  must  be  designed  uath  particular 
reference  to  minimizing  this  force,  and  our  starting-point  must  be  a  study  of 
friction. 


BEARINGS  AND  THEIR  LUBRICATION 

Bearing  Surfaces 
Cylindrical. 
Journal. . .   \    Conical. 
Spherical. 


Bearings  with  sliding  contact. 


Thrust 


Flat. 

Conical. 

Spherical. 

Generated  by  curve  of  tractrix. 


Bearings  with 
rolling  contact. 


Ball  bearings. 


Roller  bearings. 


f  Radial. 


Thrust. 


Journal. . 


Thrust , 


Spherical  (of  balls)   with  double  curved 

surfaces  (of  races). 
Spherical  (of  balls)  with  conical  (of  races) . 

Spherical  (of  balls)  with  flat  (of  seats). 
Spherical   (of  balls)  with  double  curved 
(of  seats). 

Cylindrical  with  cylindrical. 
Conical  with  conical. 

Cylindrical  with  flat. 
Conical  with  conical. 


The  cross-section,  perpendicular  to  the  axis  of  rotation,  of  all  these  bearing  surfaces  is 
necessarily  circular. 


PARTI 

BEARINGS  WITH  SLIDING  CONTACT 


SECTION  I 
FRICTION  OF  BEARINGS  WITH  SLIDING  CONTACT 

All  sliding  surfaces,  no  matter  how  carefully  they  have  been  prepared,  are 
known  to  consist  of  minute  humps  and  hollows;  this  is  true  even  of  the  smooth- 
est surfaces  that  can  be  made  although,  of  course,  the  height  of  the  depressions 
and  hollows  varies  with  different  kinds  of  materials  and  different  degrees  of 
finish.  This  is  similar  to,  though  much  less  in  degree  than  the  ''nap  "  of  woolen 
cloth  and  the  "pile "  of  velvet.  It  may  also  be  likened  in  an  exaggerated  degree 
to  the  "fur"  on  a  rough  planed  or  milled  surface  of  cast  iron. 

When  two  solid  surfaces  are  held  in  contact  by  an  appreciable  force,  these 
minute  parts  of  the  surfaces  interlock  and  resist  relative  motion.  This  resist- 
ing force  acts  tangentially  to  the  surface  of  separation,  and  is  the  force  of 
friction. 

In  addition  to  this  interlocking  action  of  surfaces,  there  is  another  that  takes 
place  between  surfaces  that  are  very  carefully  fitted  together  and  are  in  intimate 
contact.  This  action  is  called  adhesion  and  still  further  resists  relative  motion. 
One  of  the  best  examples  of  this  adhesion  of  accurately  finished  surfaces  is  the 
action  of  the  Swedish  gages  that  are  now  familiar  to  most  engineers  and 
machinists.  The  working  surfaces  of  these  gages  have  a  splendid  lapped  finish. 
When  two  of  these  surfaces  are  rubbed  into  intimate  contact  they  will  adhere 
with  a  force  having  an  intensity  several  times  that  of  atmospheric  pressure;  that 
is,  the  force  of  adhesion  between  the  surfaces  tending  to  prevent  separation  by 
a  direct  pull  is  much  greater  than  the  force  exerted  by  the  pressure  of  the  atmos- 
phere upon  the  exposed  surfaces.  Experiments  witnessed  by  the  author  showed 
that  the  force  required  to  slide  one  of  the  gages  upon  another  was  about  21/2 
lb.  when  the  two  were  in  intimate  contact.  The  surfaces  were  i  3/8  in.  long 
by  3/8  in.  wide,  representing  an  area  of  about  1/2  sq.  in.  The  upper  gage 
was  1/2  in.  thick,  thus  giving  a  weight  of  about  i  oz.  This  force  of  adhesion 
is  but  little  understood,  but  this  simple  experiment  shows  how  much  it  can  be 
quantitatively  between  accurate  surfaces. 

Just  as  the  intimate  contact  of  the  surfaces  tends  to  increase  the  resistance  of 
the  relative  motion,  a  separation  of  these  surfaces  by  oil,  grease  or  other  lubri- 
cant reduces  this  resistance  to  a  considerable  extent.  This  fact  will  lead  us  later 
on  to  a  consideration  of  the  lubrication  of  bearings. 

In  some  machine  elements,  as  clutches  and  brakes,  this  resistance  to  relative 

5 


6  BEARINGS  AND  THEIR  LUBRICATION 

motion  is  a  desiraoie  thing;  but  in  bearings,  on  the  contrary,  it  is  undesirable 
as  it  requires  power  to  overcome  it  and,  if  excessive,  may  cause  the  surfaces  to 
cut  and  ruin  the  bearings.  Thus  a  knowledge  of  the  laws  of  friction  is  abso- 
lutely essential  to  the  design  of  successful  bearings. 

Unfortunately,  these  laws  are  not  thoroughly  understood,  and  it  is  now 
recognized  that  many  of  the  older  and  accepted  principles  only  hold  true  for  a 
very  limited  range  of  conditions. 

FRICTION  OF  REST  AND  MOTION 

Experiment  has  shown  that  it  requires  a  greater  force  to  start  one  surface  to 
sliding  over  another  than  to  continue  this  sliding  after  motion  has  begun. 
Therefore,  two  forces  of  friction  are  recognized;  the  friction  of  rest  (sometimes 
called  stiction)  and  the  friction  of  motion.  A  knowledge  of  the  laws  of  the 
friction  of  motion  is  of  more  importance  and  for  that  reason  the  friction  of 
rest  will  be  treated  briefly. 

Again,  the  laws  of  friction  naturally  divide  themselves  into  two  groups 
referred  to  the  condition  of  the  surfaces  in  contact;  (a)  for  dry  or  unlubricated 
surfaces,  (b)  for  lubricated  surfaces. 

COEFFICIENT  OF  FRICTION 

In  a  preceding  paragraph  it  is  stated  that  the  force  of  friction  is  the  re- 
sisting force  to  relative  motion  when  two  solid  surfaces  are  held  in  contact  by  an 
appreciable  force.  The  ratio  of  the  force  required  to  produce  motion  to  the 
force  holding  the  surfaces  in  contact  is  called  the  coefficient  of  friction. 

For  flat  surfaces,  if  we  let  F  equal  the  frictional  resistance,  \V  the  normal 
load,  and  /  the  coefficient  of  friction,  then 

F=fW,  or  /=  -. 

For  flat  surfaces  where  the  top  body  is  acted  upon  by  its  own  weight  only, 
the  coefficient  of  friction  is  equal  to  the  tangent  of  the  angle  of  repose;  that  is, 
the  angle  of  inclination  with  the  horizontal  of  the  plane  of  contact  of  the  two 
surfaces  where  sliding  will  begin  to  take  place.  If  this  angle  is  indicated  by  a, 
the  coefficient  of  friction  equals  the  tangent  of  angle  a.     Or,  /=tan.  a. 

In  a  similar  manner  the  coefficient  of  friction  for  cylindrical  bearings  is  the 
ratio  of  the  frictional  resistance  to  the  normalpressure  and  is  again  indicated  by 
/.  The  character  and  fit  of  the  surfaces,  however,  have  a  great  deal  to  do  with 
the  value  of  the  coefficient  for  such  cases;  the  distribution  of  the  normal  pres- 
sure is  variable,  and  the  intensity  of  the  normal  pressure  is  a  difficult  quantity  to 


SLIDING  FRICTION  7 

determine  accurately.  For  convenience,  it  is  customary  to  take  as  the  normal 
pressure  the  intensity  of  pressure  per  unit  of  projected  area  of  the  bearing.  Thus, 
if  d  is  the  diameter  of  the  shaft  and  /  its  length,  the  intensity  of  pressure  per  unit 
of  projected  area  is 

W 

1)0=  ---. 
dl 

The  reason  for  the  difficulty  in  determining  accurately  the  intensity  of  normal 
pressure  is  evident  to  anyone  familiar  with  bearings.  It  is  perfectly  possible  to 
fit  a  journal  into  its  bearing  so  that  there  shall  be  a  complete  arc  of  contact  for 
1 80  degrees  of  bearing  circumference.  In  practice,  however,  a  journal  is  made 
free  in  its  bearing,  the  amount  of  freeing  depending  upon  circumstances.  Thus 
the  arc  of  contact  is  decreased.  In  many  cases  this  arc  may  not  be  over  one- 
half  of  the  total  bearing  depth  of  the  journal,  or  say  about  120  degrees.  Again, 
journals  and  bearings  tend  to  wear  in  use  and  thus  produce  a  varying  amount 
of  contact  depending  upon  the  amount  and  location  of  the  wear  that  has  taken 
place. 

WORK  OF  FRICTION 

The  work  of  friction  Is  an  important  point  to  consider  in  the  design  of  bear- 
ings, for  it  gives  a  measure  of  the  amount  of  energy  transformed  into  heat  by  the 
frictional  resistance.  This  quantity  of  heat  has  a  direct  bearing  upon  the  per- 
formance and  life  of  many  types  of  bearings.  In  some  cases,  a  circulation  of 
water  or  oil  is  established  to  conduct  away  the  heat  thus  liberated. 

For  flat  plates  H  the  foot  pounds  of  energy  absorbed  per  minute  equals 
/  W  V,  where  v  is  the  velocity  in  feet  per  minute  and  W  is  expressed  in  pounds. 
For  circular  surfaces,  if  n  is  the  number  of  revolutions  per  minute 

frrdnW 
H=-  =o.26iSfdnW.   ■ 

12 

Table  i  gives  the  moment  of  friction  in  inch  pounds  and  the  energy  lost 
in  friction  in  foot  pounds  per  minute  for  the  ordinary  types  of  machinery  bear- 
ing surfaces.     The  notation  is  as  follows: 

/     =  Coefficient  of  friction. 

W  =load  on  journals  or  pivots  in  pounds. 

r     =  radius  in  inches. 

r^  =  inner  radius  of  collar  in  inches. 

^2  =  outer  radius  of  collar  in  inches. 

d    =  diameter  in  inches. 

n    =  number  of  revolutions  per  minute. 

b    =  half  of  the  included  angle  of  cone. 


BEARINGS  AND  THEIR  LUBRICATION 


Kind  of  bearing  surfaces 


Moment  of  friction  in 
inch  pounds 


Energy  lost  in  friction 

in  foot  pounds  per 

minute 


Journal  bearings  as  ordinarily  fitted 

0.5/^1^ 

0. 2618  fdWn. 

Journal  bearings  tightly  fitted . . 

0.62s  fdW 

0. 3:^  24  fdWn. 

Journal  bearings  fitted  to  grasp  the  shaft 
and  give  a  uniform  pressure  throughout. 

0.78/^t^ 

0.4112  fdlVn. 

Flat  pivot  bearings 

0.66  frW 

0 .  T,4g  fWrn. 

Flat  collar  bearings 

"-KvIt;: 

0.349/^l^w^; 

Conical  pivot  bearings 

0.66  fWr  cosec.  b 

0 .  T,4C)  fWrn  cosec.  b. 

Conical    journal     bearings     or 
journal  bearings. 

tapered 

0. 66 fWr  sec.  b 

0.349/I'Fm  sec.  b. 

Truncated  cone  pivot  bearings 

J,    3_^    3 

o.66fW  '        ' 

^2  sin.  b 

m/   '"-'^^ 

0.349/PFw        . 

^2  sm.  b 

Hemispherical  pivot  bearings. . . 

frW           

o.26i8fdWn. 

(tractrix 

frW 

0. 2618  fdWn. 

curve). 

Table  i. — Moment  of  Friction  and  Work  of  Friction  for  Bearings  with  Sliding 

Surfaces. 

The  mathematical  work  showing  the  derivation  of  these  formulas  of  Table  i 
can  be  found  in  Thurston's  Friction  and  Lost  Work,  beginning  on  page  40. 

Attention  should  be  directed  to  the  three  items  in  Table  i  dealing  with 
journal  bearings.  The  first  is  for  bearings  as  ordinarily  fitted;  the  second  with 
bearings  tightly  fitted  and  shows  that  the  loss  of  energy  in  friction  is  1.27  times 
that  of  properly  fitted  bearings.  The  third  deals  with  bearings  that  have  been 
so  fitted,  or  have  so  worn  that  the  intensity  of  pressure  is  uniform  throughout. 
Such  a  condition  produces  a  loss  of  power  1.57  times  that  of  properly  fitted 
bearings. 

POWER  LOST  IN  BEARINGS 

To  obtain  the  horsepower  lost  in  bearings,  or  the  amount  of  power  required 
to  drive  a  journal  or  pivot  against  its  own  friction  divide  the  quantities  obtained 


SLIDING  FRICTION  9 

from  the  formulas  in  the  last  column  of  Table  i  by  33,000.  Similarly,  to  obtain 
the  number  of  British  thermal  units  of  heat  liberated  in  a  given  bearing,  divide 
the  quantities  obtained  from  the  formulas  in  the  last  column  of  Table  i  by  778. 

FRICTION  OF  UNLUBRICATED  SURFACES 

The  early  experimental  work  to  determine  the  values  of  the  coefficient  of 
friction  for  various  surfaces  in  contact  v^as  done  by  Coulomb,  Morin  and 
Rennie.  The  work  of  General  Morin,  a  most  original  and  scientific  engineer, 
was  the  most  extensive.     From  his  experiments  he  deduced  three  laws,  which  are 

1.  The  frictional  resistance  is  proportional  to  the  pressure. 

2.  The  frictional  resistance  is  independent  of  the  speed. 

3.  The  frictional  resistance  is  independent  of  the  extent  of  the  surfaces.^ 

These  laws  have  frequently  been  quoted  and  treated  as  if  they  were  rigidly 
true.  Such,  however,  is  not  the  case.  They  are  now  known  to  be  only  ap- 
proximately true  for  a  very  limited  range  of  conditions.  In  fact,  General 
Morin  himself  considered  them  only  as  approximations  as  shown  by  this  para- 
graph from  a  letter  written  by  him  to  the  secretary  of  the  Institute  of  Mechanical 
Engineers,  England,  and  printed  in  the  proceedings  of  that  Institution  for  the 
year  of  1883,  at  page  663.  "The  results  furnished  by  my  experiments  as  to  the 
relations  between  pressure,  surface,  and  speed,  on  the  one  hand,  and  sliding 
friction  on  the  other,  have  always  been  regarded  by  myself,  not  as  mathematical 
laws,  but  as  the  close  approximations  to  the  truth  vdthin  the  limits  of  the  data 
of  the  experiments  themselves.  The  same  holds,  in  my  opinion,  for  many  other 
laws  of  practical  mechanics;  such  as  those  of  rolling  resistance,  etc." 

Table  2  is  translated  from  the  work  of  General  Morin  and  gives  the  results 
of  some  of  his  experiments  on  plain  surfaces.  The  values  are  for  static  friction, 
or  friction  at  very  low  velocity,  with  light  pressures.  The  surfaces  were  dry,  or 
but  very  slightly  lubricated. 

In  Table  2  it  will  be  noticed  that  the  values  in  the  last  column  are  consider- 
ably greater  than  those  in  the  first.  This  shows  us  the  effect  of  leaving  compara- 
tively soft  bodies  in  contact  for  a  considerable  length  of  time.  The  irregular 
particles  of  the  surfaces  tend  to  compress  and  interlock  to  a  greater  degree  than 
if  the  surfaces  are  merely  brought  in  contact  and  motion  at  once  takes  place. 

The  experiments  of  Rennie  in  1829  were  made  in  general  with  dry  and  un- 
lubricated  surfaces.  The  principles  that  he  deduced  were  that  the  friction  of 
sliding  surfaces  differs  with  the  character  of  the  surfaces;  that  the  friction  of 
lubricated  surfaces  depends  upon  the  lubricant,  rather  than  upon  the  bodies 
themselves;  that  friction  is  least  with  hard  materials  and  greatest  with  soft  ones; 

^See  Annales  des  Mines,  3  serie,  October,  1836,  p.  27. 


lO 


BEARINGS  AND  THEIR  LUBRICATION 


Mutual  arrange- 
ment of  the  fibers 

Coefficient  of  friction 

Surfaces  in  contact 

When  the 

body  is  in 

motion 

When  the  surfaces 
have    been   some 
time  in  contact 

Oak  on  oak,  dry 

Oak  on  oak,  dry 

Oak  on  oak,  wet 

Elm  on  oak,  dry 

Elm  on  oak,  dry 

Ash  on  oak,  dry 

Fir  on  oak,  dry 

Beech  on  oak,  dry 

Wild  pear  tree  on  oak,  dry 

Wrought  iron  on  oak,  dry 

Yellow  copper  on  oak,  dry 

Black  curried  leather  on  oak,  dry. .  . 
Cowhide  sole  leather  laid  flat  on  oak,  dry. 
Sole  leather  on  edge  on  oak,  dry. .  .  . 
Sole  leather  on  edge  on  oak,  wet. . .  . 
Mat  of  small  hempen  cords  on  oak,  dry. 

Parallel 

Perpendicular .  . 
Perpendicular .  .  . 

Parallel 

Perpendicular . .  . 

Parallel 

Parallel 

Parallel 

Parallel 

Parallel 

Parallel 

Parallel 

Parallel 

Parallel 

Parallel 

Parallel 

G.48 
G.32 
G.25 
0.43 
0.45 
G.40 
0.36 
0.36 
0.4G 
G.62 
G.62 
G.27 
G.52 
0.34 

G.  29 
0.32 

o.6Ga  G.65 

0-54 
G.71 
0.69 
0-57 
0.50 
G.52 

0-53 
0.44 
G.62 
G.62 

0.74 
G.61 

0-43 
G.79 
0.50 

Table  2. — Coefficients  of  Friction  for  Various  Surfaces  in  Contact  {Morin). 

that  the  limit  of  abrasion  of  two  rubbing  surfaces  is  determined  by  the  hardness 
of  the  softer,  and  that  with  wood  and  metals  friction  varies  with  the  pressure 
and  is  independent  of  the  extent  of  the  surface,  time  of  contact  and  velocity. 
Table  3  is  from  Rennie's  experiments. 


Pressure  in  pounds 
per  square  inch 

Brass  on 

cast  iron 

Wrought  iron  on 
wrought  iron 

Wrought  iron  on 
cast  iron 

Steel  on  cast 
iron 

187 

0.23 
G.22 
0.21 
G.21 
G.23 
G.23 
G.23 

0.25 
0.27 

0.31 

0.38 

0.41 

abraded 

abraded 

G.28 
G.29 
0-33 
0-37 
0-37 
0.38 
abraded 

0.30 
0-33 
0-35 
0.35 
G.36 

224 

336 

448 

Kdo    . 

672 

784 

abraded 

Table  3. — CoEFFfciENXs  of  Friction  of  Unlubricated  Surfaces  {Rmnie). 


SLIDING  FRICTION 


II 


Tables  4  to  7,  inclusive,  are  also  from  General  Morin's  experiments  and  were 
made  at  low  velocities,  with  surface  either  dry,  oily,  or  greasy,  as  stated  in  the 
tables.  In  the  original  table  of  General  Morin's  from  which  Table  5  has  been 
translated,  there  is  a  column  giving  the  position  of  the  fibres  of  the  wooden  test 
pieces  with  reference  to  the  direction  of  motion,  and  in  some  cases  with  refer- 
ence to  the  supposed  direction  of  the  fibres  of  the  metal  test  pieces.  How- 
ever, the  references  are  not  entirely  clear  and  have  been  omitted  in  the  trans- 
lations; apparently  all  of  the  fibres  of  the  wooden  test  pieces  ran  parallel  with 
the  direction  of  motion  and  parallel  with  the  supposed  direction  of  the  fibers 
of  the  metal  pieces.  Similarly  a  column  has  been  dropped  in  translating 
Table  6,  with  the  belief  that  the  values  are  not  rendered  less  useful  thereby. 
Inconsistencies  that  may  be  noted  in  the  values  of  the  coefficients  of  friction 
may  be  explained  by  variations  in  the  conditions  of  the  tests. 


Surfaces  in 
contact 


Condition  of  the  surfaces 


Mutual  arrangement  of 
the  fibers 


Oak  on  oak '  Coated  with  dry  soap '  Parallel. 

Oak  on  oak Coated  with  tallow Parallel. 

Oak  on  oak Coated  with  lard I  Parallel. 

Oak  on  oak Oily 

Oak  on  oak Dry 

Oak  on  oak Coated  with  tallow. 

Oak  on  oak Coated  with  lard. . . 

Oak  on  oak Oily 

Beech  on  oak. ...  Coated  with  tallow. 


Oily. 


Beech  on  oak. 

Elm  on  oak Coated  with  dry  soap. 

Elm  on  oak Coated  with  tallow. . . 

Elm  on  oak. 
Elm  on  oak. 
Elm  on  elm. 
Elm  on  elm. 


Coated  with  lard 

Oily 

Coated  with  dry  soap 

Oily 

Oak  on  elm Dry 

Oak  on  elm Coated  with  dry  soap 

Oak  on  elm Coated  with  tallow Parallel 

Oak  on  elm Coated  with  lard |  Parallel 

Oak  on  elm Oily Parallel 


Parallel 

Perpendicular . 
Perpendicular . 
Perpendicular . 
Perpendicular . 

Parallel 

Parallel 

Parallel 

Parallel 

Parallel 

Parallel 

Parallel 

Parallel 

Parallel 

Parallel 


Coefficient  of 
friction 


o.  164 
0.075 
0.067 
0.108 
0.336 
0.083 
0.072 
0.143 
0-055 
0153 
0.137 
0.070 
0.060 
o.  119 
0.139 
0.140 
0.246 
0.136 
0.073 
0.066 
0.136 


Table  4. — Coefficients  of  Friction  for  Wood  on  Wood,  Dry  or  Scantily  Lubricated 

Surfaces  {Morin). 


12 


BEARINGS  AND  THEIR  LUBRICATION 


Surfaces  in  contact 


Condition  of  the  surfaces 


Coefficient  of 
friction 


Wrought  iron  on  oak 
Wrought  iron  on  oak 
Wrought  iron  on  oak 
Cast  iron  on  oak  .... 

Cast  iron  on  oak 

Cast  iron  on  oak .... 

Cast  iron  on  oak 

Cast  iron  on  oak 

Cast  iron  on  oak 

Cast  iron  on  oak 

Copper  on  oak 

Copper  on  oak 

Cast  iron  on  elm .... 
Cast  iron  on  elm  .... 
Cast  iron  on  elm  .... 
Cast  iron  on  elm  .... 

Cast  iron  on  elm 

Cast  Iron  on  elm 


Moistened  with  water. 
Coated  with  dry  soap . . 
Coated  with  tallow. . . . 
Dry 


Coated  with  dry  soap  . 
Moistened  with  water. 
Coated  with  tallow  .  .  . 

Coated  with  lard 

Coated  with  olive  oil.  , 

Oily 

Coated  with  tallow  . 

Oily 

Dry 


Wrought  iron  on  elm 

Wrought  iron  on  elm 

Wrought  iron  on  elm 

Wrought  iron  on  elm 

Wrought  iron  on  elm  .... 

Oak  on  cast  iron 

Oak  on  cast  iron 

Elm  on  cast  iron 

Elm  on  cast  iron 

Lignum-vitae  on  cast  iron. 
Lignum-vitae  on  cast  iron. 
Lignum-vitae  on  cast  iron. 

Oak  on  wrought  iron 

Oak  on  wrought  iron 

Lignum-vitae  on  bronze.  .  , 
Lignum-vitae  on  bronze.  .  , 
Lignum-vitae  on  bronze.  . 


Coated  with  tallow 

Coated  with  olive  oil 

Coated  with  lard  and  plumbago.  .  . 
Greasy  after  coating  with  tallow.  .  .. 
Greasy  after  coating  with  lard  and 

plumbago. 
Dry 


Coated  with  tallow.  . 
Coated  with  lard .... 
Coated  with  olive  oil . 

Oily 

Coated  with  tallow.  . 

Oily 

Coated  with  tallow .  . 

Oily 

Coated  with  tallow. . . 
Coated  with  olive  oil. 

Oily 

Coated  with  tallow . . 

Oily 

Coated  with  tallow .  . 
Coated  with  olive  oil. 
Oilv 


0.256 
o.  214 
0.085 
0.490 
0.189 
0.218 
0.078 
0.075 
0.075 
0.107 
0.069 
o.ioo 

0.195 
0.077 

0.061 
0.091 

0.125 
0.137 

0.252 
0.078 
0.076 

0-055 
0.138 

0.080 

0.168 

0.066 

0-135 
0.074 

0.076 

O.I2I 
0.098 
0.149 
0.082 

0-053 
0.146 


Table  5. — Coefficients  of  Friction  for  Wood  and  Metal,  Dry  or  Scantily 
Lubricated  Surfaces  (Morin). 


SLIDING  FRICTION 


13 


Surfaces  in  contact 


Condition  of  the  surfaces 


Cast  iron  on  cast  iron 

Cast  iron  on  cast  iron 

Cast  iron  on  cast  iron 

Cast  iron  on  cast  iron 

Cast  iron  on  cast  iron 

Cast  iron  on  cast  iron 

Cast  iron  on  cast  iron 

Cast  iron  on  cast  iron 

Wrought  iron  on  cast  iron . . . . 
Wrought  iron  on  cast  iron. . . . 
Wrought  iron  on  cast  iron. . . . 

Wrought  iron  on  cast  iron 

Steel  on  cast  iron 

Steel  on  cast  iron 

Steel  on  cast  iron 

Steel  on  cast  iron 

Steel  on  cast  iron 

Yellow  copper  on  cast  iron .  . . 
Yellow  copper  on  cast  iron. . . 
Yellow  copper  on  cast  iron . . . 
Yellow  copper  on  cast  iron . . . 
Yellow  copper  on  cast  iron. . . 

Bronze  on  cast  iron 

Bronze  on  cast  iron 

Bronze  on  cast  iron 

Bronze  on  cast  iron 

Cast  iron  on  wrought  iron. . . . 
Cast  iron  on  wrought  iron. . . . 
Cast  iron  on  wrought  iron .... 
Wrought  iron  on  wrought  iron 
Wrought  iron  on  wrought  iron 
Wrought  iron  on  wrought  iron 
Wrought  iron  on  wrought  iron 
Wrought  iron  on  wrought  iron 

Steel  on  wrought  iron 

Steel  on  wrought  iron 

Bronze  on  wrought  iron 

Bronze  on  wrought  iron 

Bronze  on  wrought  iron 

Bronze  on  wrought  iron 

Bronze  on  wrought  iron 

Cast  iron  on  bronze 

Cast  iron  on  bronze 


Dry 

Moistened  with  water 

Coated  with  soap 

Coated  with  tallow 

Coated  with  lard 

Coated  with  olive  oil 

Coated  with  lard  and  plumbago. 

Oily 

Dry 

Coated  with  tallow , 

Coated  with  lard , 

Coated  on  olive  oil 

Dry 


Coated  with  tallow. 
Coated  with  lard. . . 
Coated  on  olive  oil . 

Oily 

Dry 


Coated  with  tallow. . 
Coated  with  lard .... 
Coated  with  olive  oil , 

Oily 

Dry 


Coated  with  tallow. . 
Coated  with  olive  oil . 

Oily 

Coated  with  tallow. . 

Coated  with  lard 

Coated  with  olive  oil . 
Dry 


Coated  with  tallow . . 

Coated  with  lard 

Coated  with  olive  oil , 

Oily 

Coated  with  tallow. . 
Coated  with  lard .... 
Dry 


Coated  with  tallow 

Coated  with  lard  and  plumbago. 

Coated  with  olive  oil 

Oily 

Dry 

Coated  with  tallow 


Coeflficient  of 

friction 

0.152 

0.314 

0.197 

O.IOO 

0.070 

0.064 

0-055 

0.144 

0.194 

0.103 

0.076 

0.066 

0.202 

0.105 

0.081 

0.079 

0.109 

0.189 

0.072 

0.068 

0.066 

0.II5 

0.217 

0.086 

0.077 

0.107 

0.098 

0.053 

0.063 

0.138 

0.082 

0.081 

0.070 

0.177 

0.093 

0.076 

0.I6I 

0.081 

0.089 

0.072 

0.166 

0.147 

0.085 

Table.  6. — Coefficients  of  Friction  for  Metal  on  Metal,  Dry  or  Scantily 
Lubricated  Surfaces  {Morin).        ' 


14 


BEARINGS  AND  THEIR  LUBRICATION 


Surfaces  in  contact 


Cast  iron  on  bronze 

Cast  iron  on  bronze 

Cast  iron  on  bronze 

Wrought  iron  on  bronze 

Wrought  iron  on  bronze 

Wrought  iron  on  bronze 

Wrought  iron  on  bronze 

Wrought  iron  on  bronze 

Steel  on  bronze 

Steel  on  bronze 

Steel  on  bronze 

Steel  on  bronze 

Bronze  on  bronze 

Bronze  on  bronze |  Coated  with  ohve  oil 

Bronze  on  bronze Oily 


Condition  of  the  surfaces 


Coated  with  lard 

Coated  with  olive  oil . . 

Oily 

Dry 

Coated  with  tallow 

Coated  with  lard 

Coated  with  olive  oil 

Oily 

Dry 

Coated  with  tallow 

Coated  with  olive  oil 

Coated  with  lard  and  plumbago.. 
Dry 


Coefficient  of 
friction 


0.070 
0.067 
0.132 
0.172 
0.103 
0.075 
0.078 
0.160 
0.152 
0.056 
0.053 
0.067 
0.201 
0.058 
0134 


Table  6  (Cow^iwMsc?).— Coefficients  of   Friction  for  Metal  on  Metal, 
Scantily  Lubricated  Surfaces  {Morin). 


Dry  or 


Surfaces  in  contact 


Condition  of  the  surfaces 


Arrangement  of  the 
surfaces 


Coefficient 
of  friction 


Tanned  cowhide  on  cast     Dry j  The  leather  laid  flat. . . 

iron,  j 

Tanned  cowhide  on  cast     Moistened    a  n  d     saturated    The  leather  laid  flat. .  . 


iron. 
Tanned  cowhide  on  cast 

iron. 
Tanned  cowhide  on  cast 

iron. 
Tanned  cowhide  on  cast 

iron. 
Tanned  cowhide  on  cast 

iron. 
Tanned  cowhide  on  cast 

iron. 


with  water.  | 

Coated  with  tallow The  leather  laid  flat. . 

Coated  with  oHve  oil The  leather  laid  flat. . 


The   leather   oily,    the   cast     The  leather  laid  flat 
iron  moistened  with  water 
Moistened  with  water 


0.559 
0.365 
0.159 

0.133 
0.229 


Moistened  with  olive  oil. 


The     leather     placed         0.338 

edgewise. 
The     leather     placed         0.135 

edgewise. 


Table  7. — Coefficients  of  Friction  for  Leather  on  Cast  Iron,  Dry  or  Scantily 
Lubricated  Surfaces  {Morin). 

As  speeds  increase  the  value  of  the  coefficient  of  friction  decreases,  but  the 
amount  of  this  decrease  is  unknown  experimentally,  except  for  certain  com- 
binations of  metals.     In  Elements  of  Machine  Design  by  Kimball  and  Barr, 


SLIDING  FRICTION  1 5 

page  99,  the  coefficient  of  friction  for  cast  iron  or  steel  at  a  velocity  of  440  ft. 
per  minute  is  given  as  0.32.  At  2460  ft.  per  minute  as  0.2,  and  at  5280  ft.,  or 
a  mile  a  minute,  as  0.006.  It  is  also  stated  that  these  data  for  cast  iron  on  steel 
will  serve  as  a  rough  guide  to  what  may  be  expected  to  occur  for  other  com- 
bination of  materials. 

A  consideration  of  the  values  of  the  coefficient  of  friction  in  Table  2  show 
that  a  great  decrease  takes  place  as  soon  as  motion  begins,  even  although  the 
velocity  is  very  low.  It  might  be  supposed  that  the  change  is  very  abrupt  from 
t"he  friction  of  rest  of  the  friction  of  motion,  but  it  is  now  generally  believed  that 
this  change  is  gradual,  and  that  the  value  of  the  coefficient  at  rest  is  not  far 
different  from  the  value  at  a  very  slow  speed. 

LAWS  OF  DRY  OR  UNLUBRICATED  FRICTION 

John  Goodman  sums  up  the  laws  governing  dry  friction  in  this  manner, 
Proc.  Inst.  C.  E.,  Vol.  LXXXV,  page  391. 

1.  The  friction  between  dry  surfaces  under  moderate  loads  and  low  veloci- 
ties varies  directly  as  the  normal  pressure  between  them. 

2.  The  normal  pressure  remaining  unchanged,  the  friction  is  independent 
of  the  area  in  contact. 

3.  The  friction  is  always  greater  on  the  reversal  of  direction  of  sliding. 

4.  The  friction  sensibly  diminishes  with  a  rise  of  the  temperature. 

These  laws  do  not  differ  materially  from  those  of  General  Morin,  except 
that  they  are  restricted  to  low  velocities  and  moderate  loads.  As  a  matter  of 
fact,  these  were  the  conditions  under  which  General  Morin  experimented. 

FRICTION  OF  LUBRICATED  SURFACES 

The  laws  for  lubricated  friction  must  be  considered  from  an  entirely  different 
standpoint  than  those  from  dry  friction,  although  it  is  very  difficult,  if  not 
impossible,  to  define  accurately  where  one  condition  begins  and  the  other  ends. 
In  a  journal  bearing,  properly  lubricated,  the  friction  resistance  must  obviously 
conform  to  the  laws  for  lubricated  surfaces.  On  the  other  hand,  if  the  supply  of 
lubricant  is  intermittent,  as  the  quantity  diminishes  the  resistance  will  tend  to 
conform  to  the  laws  for  dry  or  unlubricated  surfaces.  This  latter  condition  is 
represented  by  the  usual  oil  hole  and  squirt-can  method  of  oiling.  Other  fami- 
liar forms  of  oiling  are  by  means  of  sight  feed  and  syphon  lubricators,  pad  lubri- 
cators, oiling  rings  and  chains,  and  oil  baths.  The  efficiency  of  these  devices 
increases  in  the  order  named. 

Lubricant  is  introduced  between  bearing  surfaces  in  order  to  separate  them 
by  a  film  of  lubricant  and  thereby  change  the  frictional  resistance  from  that 


1 6  BEARINGS  AND  THEIR  LUBRICATION 

of  the  solid  surfaces  in  contact  to  the  fluid- resistance  of  the  film.  Thus  the 
nature  and  condition  of  the  lubricant  has  more  to  do  with  this  kind  of  friction 
than  the  bearing  surfaces  themselves  assuming  that  they  are  well  fitted. 

FRICTION  OF  REST  OF  LUBRICATED  SURFACES 

It  is  frequently  observed  that  after  a  machine  has  been  idle  for  some  time, 
it  ''starts  hard."  The  bearings  are  sometimes  said  to  be  ''stiff."  The  reason 
for  this  is  that  the  pressure  of  the  journals  in  the  bearings  tends  to  squeeze  out 
the  lubricant  and  allows  the  metallic  surfaces  to  come  into  more  or  less  close 
contact.  Experience  has  shown  that  it  is  very  difficult  even  with  small  areas 
and  heavy  loads  to  squeeze  out  all  the  lubricant,  but  its  thickness  will  be  very 
much  reduced  if  the  bearing  is  idle  for  any  considerable  length  of  time.  Thus 
when  a  machine  is  started  after  a  long  shut  down  the  frictional  resistance  in  its 
bearings  is  more  nearly  that  of  dry  or  unlubricated  surfaces  than  of  well  lu"bri- 
cated  surfaces.  This  fact  shows  the  necessity  of  carefully  oiling  a  heavy  machine 
before  it  is  started.  If  this  is  not  done  the  surface  of  journals  and  bearings  may 
be  in  such  close  contact  that  abrasion  will  take  place  before  the  lubricant  re- 
maining in  the  bearing  can  be  distributed  and  an  oil  film  established. 

The  coefficients  of  rest  or  starting  of  motion  as  given  by  Professor  Thurston 
in  Friction  and  Lost  Work,  page  319,  for  sperm  oil,  lard  oil  and  mineral  oil  are, 
respectively:  0.15;  o.i;  0.15.  These  are  very  much  higher  than  similar  coeffi- 
cients for  well-lubricated  surfaces  when  in  motion.  These  values  were  obtained 
from  experiments  made  on  a  testing  machine  at  pressures  ranging  from  75  to 
500  lb.  per  square  inch  and  with  lubricant  between  metallic  surfaces. 

FRICTION  OF  POORLY  LUBRICATED  SURFACES 

In  the  preceding  paragraph  is  pointed  out  the  fact  that  the  friction  of  lubri- 
cated surfaces  at  rest  is  much  greater  than  the  friction  of  similar  surfaces  well 
lubricated  when  in  motion.  Between  these  two  extremes  is  the  condition  of 
imperfect  lubrication.  This  is  the  condition  under  which  most  machine  bear- 
ings are  run.  Yet  there  are  so  many  variables  that  it  is  impossible  to  establish 
any  laws.  The  better  the  imperfect  lubrication,  the  more  nearly  will  the  condi- 
tions conform  to  the  laws  for  well  lubricated  surfaces;  The  poorer  the  lubrica- 
tion, the  more  nearly  will  these  conditions  conform  to  the  laws  of  dry  surfaces. 

The  variables  that  govern  the  condition  of  poorly  lubricated  surfaces  are  the 
character  of  the  surfaces  themselves,  their  fit,  the  supply  and  nature  of  the  lubri- 
cant, the  pressure  between  them,  the  velocity  of  rubbing  and  the  temperature 
as  affecting  the  viscosity  of  the  oil  or  grease.  With  so  many  variables  it  is  easy 
to  see  the  difficulties  of  obtaining  results  from  even  the  most  careful  experi- 
mental work. 


SLIDING  FRICTION  17 

Efforts  have  been  made  to  state  the  laws  of  imperfect  friction,  but  with  very 
little  success.  It  is  probably  of  more  value  to  consider  briefly  some  of  the 
actions  that  take  place  in  a  bearing. 

The  pressure  tends  to  squeeze  out  the  lubricant.  If  this  pressure  is  very 
high  it  may  be  impossible  to  get  a  sufficient  supply  of  oil  between  the  surfaces  to 
produce  a  proper  film.  Then  the  bearing  would  be  practically  in  metallic 
contact  with  its  journal,  and  the  coefficient  of  friction  would  be  nearly  that  of 
dry  surfaces  in  contact. 

Again,  if  the  velocity  of  rubbing  is  so  very  low  that  the  lubricant  cannot  be 
drawn  between  the  surfaces  to  form  a  film,  there  may  be  practically  metallic 
contact. 

This  same  condition  may  be  produced  by  an  insufficient  quantity  of  oil  and 
this  is  the  most  general  cause  of  imperfect  lubrication.  If  the  pressure  and 
velocity  of  rubbing  are  normal  the  supply  of  oil  or  grease  may  be  such  as  to  give 
any  condition  of  lubrication  between  that  of  unlubricated  surfaces  and  well 
lubricated  surfaces. 

The  influence  of  temperature  is  important,  for  even  with  a  suitable  supply 
of  lubricant  at  a  given  pressure,  velocity  of  rubbing  and  ordinary  temperature, 
the  thinning  of  the  oil  as  the  temperature  increases  may  be  sufficient  to  break 
down  the  film  and  produce  metallic  contact. 

Excellent  lubrication  can  only  be  obtained  by  a  plentiful  supply  of  oil,  as  by 
an  oil  bath,  flooded  lubrication,  and  the  best  arrangements  of  oiling  rings.  The 
greater  number  of  machine  bearings  are  oiled  either  by  a  squirt  can  through  an 
oil  hole,  by  sight-feed  oil  cups,  syphon  oil  cups,  pads,  or  compression  grease 
cups.  All  of  these  devices  as  ordinarily  used  can  be  considered  as  producing 
nothing  but  imperfect  lubrication,  and  thereby  bear  out  the  general  statement 
that  the  great  mass  of  machinery  bearings  are  in  this  second  indefinite  condi- 
tion of  poorly  lubricated  surfaces. 

The  only  way  in  which  this  subject  can  be  satisfactorily  considered  is  by 
studying  its  limiting  factors.  In  a  former  paragraph,  page  15,  is  discussed 
the  friction  of  dry  surfaces  and  the  following  treats  of  the  friction  of  well  lubri- 
cated surfaces. 

WELL  LUBRICATED  SURFACES 

The  best  condition  for  a  bearing  to  be  in,  and  unfortunately  the  rarest,  is, 
of  course,  well  lubricated.  This  means  a  plentiful  supply  of  lubricant,  as  from 
an  oil  bath.  When  this  condition  prevails  the  surfaces  of  the  bearings  are  defi- 
nitely separated  by  an  oil  film  and  the  frictional  resistance  is  the  resistance  to 
shearing  of  this  film  due  to  the  rubbing  action.  The  journal  floats  on  the 
lubricant. 

As  the  pressure  on  the  journal  under  ordinary  conditions  has  a  definite 


l8  BEARINGS  AND  THEIR  LUBRICATION 

direction,  the  space  between  the  journal  and  bearing  is  less  at  the  point  of  nearest 
approach  than  at  a  point  diametrically  opposite.  This  means  that  the  space 
between  the  journal  and  bearing  is  a  wedge-shaped  ring.  The  action  of  the 
journal  as  it  rotates  is  to  take  oil  from  the  region  of  least  pressure  and  carry  it 
into  the  region  of  greater  pressure.  This  action  is  sufficient  to  suck  oil  against 
small  heads.  This  continual  drawing  in  of  oil  preserves  the  film  and  at  the 
same  time  necessitates  the  doing  of  a  certain  amount  of  work  in  transferring 
the  oil  from  one  point  to  another  in  the  bearing  and  in  overcoming  its  own  fluid 
resistance. 

We  are  fortunate  in  the  amount  of  experimental  work  that  has  been  done 
to  determine  the  laws  of  friction  for  well  lubricated  surfaces.  This  subject  is 
by  no  means  in  the  haze  that  surrounds  poorly  lubricated  surfaces,  although 
much  work  still  awaits  any  investigator  who  chooses  to  enter  the  field. 

THEORY  OF  BEARING  LUBRICATION 


Professor  Osborne  Reynolds  has  given  a  most  interesting  and  valuable 
theoretical  discussion  of  bearing  lubrication  in  a  paper,  ''Theory  of  Lubri- 
cation," Phil.  Trans.  (Royal  Society,  London),   1886.     This  should  be  read 
by  everyone  who  wishes  to  grasp  the  fundamentals  of  the  subject,  yet  the  empiri- 
cal rules  given  in  following  paragraphs  are  more 
directly  serviceable  for  designing  purposes. 

Professor  Kingsbury,  Journal  Am.  Soc.  Naval 
Engineers,  Vol.  IX,  No.  2,  (1897)  gives  six  laws 
for  perfect  lubrication  arranged  from  Reynold's 
paper.  The  reference  is  to  Fig.  i,  reproduced 
from  Reynolds  and  representing  a  right  section  of 
a  cylindrical  journal  and  its  brass;  the  rotation 
of  the  journal  is  counter-clockwise.  The  condi- 
tions are  such  that  the  film  is  complete. 

I.  The  vertical  components  of  the  pressure 
and  friction  balance  the  load. 

2.  The  horizontal  component  of  the  oil  pressure  balances  the  horizontal 
component  of  the  friction. 

3.  When  the  brass  is  unloaded,  its  point  of  nearest  approach  will  be  its 
middle  point.  The  pressures  are  then  symmetrically  disposed  with  reference 
to  O,  the  positive  on  the  right,  the  negati/e  on  the  left. 

4.  As  the  load  increases,  the  positive  vertical  component  must  overbalance 
the  negative  component.     This  requires  that  H  should  be  at  the  left  of  O. 

5.  As  the  load  increases,  OH  reaches  a  maximum  value  which  places  H 
nearly,  but  not  quite  as  the  left  extremity  of  the  brass,  but  still  leaves  //  small 


Fig.  I.  DIAGRAM  OF  SECTION 
OF  JOURNAL  AND  ITS  MATING 
BEARING. 


SLIDING  FRICTION  19 

as  compared  with  GH.     For  a  further  increase  of  the  load,  H  moves  back  again 
toward  O. 

6.  As  the  load  increases,  the  distance  between  the  axes  of  brass  and  journal 
increases. 

LAWS  OF  FRICTION  FOR  WELL  LUBRICATED  SURFACES 

The  following  five  empirical  laws  of  friction  for  well  lubricated  surfaces 
have  been  modified  from  four  laws  laid  down  by  John  Goodman,  Proc.  Inst. 
C.  E.,  Vol.  LXXXV,  page  380,  and  from  the  results  of  Lasche's  experiments. 
It  is,  of  course,  uncertain  just  how  far  they  apply.  But  each  law  is  amplified 
by  the  results  of  experiments. 

1 .  The  coefficient  of  friction  for  well  lubricated  surfaces  is  practically  inde- 
pendent of  the  materials  composing  the  bearings  and  journals,  but  is  dependent 
upon  the  nature  of  the  lubricant. 

2.  The  coefficient  of  friction  for  well  lubricated  surfaces  is  from  i/6toi/io 
that  for  unlubricated  or  scantily  lubricated  surfaces. 

3.  The  coefficient  of  friction  for  well  lubricated  surfaces  and  moderate 
pressures  and  speeds  varies  approximately  inversely  as  the  normal  pressure; 
that  is,  the  frictional  resistance  per  unit  of  area  is  a  constant  assuming  the 
speed  constant.     The  total  frictional  resistance  varies  directly  as  the  area. 

Or,  •J=J  ^nd  U,=f,p, 

h     Pi 

4.  The  coefficient  of  friction  for  well  lubricated  surfaces  for  a  given  pressure 
is  very  high  for  low  rubbing  velocities,  though  decreasing  rapidly  as  the  veloc- 
ity increases.  For  velocities  from  100  to  500  ft.  per  minute  it  decreases,  vary- 
ing approximately  as  the  square  roots  of  the  velocities ;  for  velocities  from  500  to 
1600  ft.  per  minute  it  decreases  at  a  lower  rate,  varying  approximately  as  the 

fifth  roots  of  the  velocities;  for  velocities  above  1600  ft.  per  minute,  it  is  practi- 
cally independent  of  the  velocity. 

Or,  for  velocities  from  100  to  500  ft.  per  minute    ^  =     -— ^ 

fi     VV2 

Or,  for  velocities  from  t;oo  to  1600  ft.  per  minute  y  =  ., — 

Or,  for  velocities  over  1600  ft.  per  minute  J^=f^ 

5.  The  coefficient  of  friction  varies  approximately  inversely  as  the  tempera- 
ture up  to  a  point  just  before  abrasion  takes  place. 

DISCUSSION   OF   THE   FIRST   LAW 

That  the  metals  forming  the  journal  and  bearing  in  well  lubricated  bearings 
have  little  influence  on  the  coefficient  of  friction  has  been  observed  by  many 


20 


BEARINGS  AND  THEIR  LUBRICATION 


investigators.  In  Traction  and  Transmission  for  January,  1903,  Lasche  describ- 
ing some  of  his  own  experiments  says: 

"The  results  of  numerous  experiments  go  to  prove  that  the  metal  forming 
the  journal — nickel  steel,  ordinary  Siemens- Martin  steel,  or  mild  steel — has  no 
marked  influence  on  the  degree  of  friction  or  friction  work." 

And  again:  "The  journal  v/as  placed  in  a  gun-metal  bushing;  in  one  with 
white-metal  lining  and  in  one  with  a  mercury  amalgam  lining.  As  in  the  case 
with  the  metal  forming  the  journal  the  various  kinds  of  bushings  gave  approxi- 
mately the  same  results." 

The  reason  for  this  is  that  the  load  is  fluid-borne  and  so  the  frictional 
resistance  is  that  of  the  oil  film  not  of  the  metallic  surfaces. 

DISCUSSION   OF   THE    SECOND    LAW 

The  truth  of  the  second  law  is  shown  by  the  following  table  taken  from  the 
work  of  Beauchamp  Tower,  Proc.  Inst.  M.  E.  for  1883,  page  651. 


Method  of  lubrication 


Pad  under  journal 

Oil  bath 

Syphon  lubricator 


Coefficient  of  friction 


Comparative  friction 


6.48 

I 
7  .06 


In  another  experiment  made  by  Mr.  Tower  with  a  load  of  293  lb.  per  square 
inch  and  a  journal  speed  of  314  ft.  per  minute  with  an  oil  bath,  a  coefficient  of 
friction  of  0.0016  was  obtained.  Under  the  same  conditions  with  a  pad  the 
coefficient  was  0.0097  which  is  six  times  the  amount  of  the  former.  Under  the 
same  condition  Mr.  Stroudley  obtained  a  coefficient  of  0.00961  which  is  in  very 
close  agreement. 


DISCUSSION   OF   THE   THIRD    LAW 

The  following  tabulation  is  from  the  experiments  of  Mr.  Tower,  Proc.  Inst. 
M.  E.  for  1883,  page  650. 

Load  in  lb.   per  sq.   in.,  520         468         415         363  310         258         205         153         100 

Frictional  resistance  per  sq.  in.,     0.416     0.514     0.498     0.472     0.464     0.438     0.43     0.458     0.45 

This  shows  that  the  unit  frictional  resistance  is  sensibly  constant  with  vary- 
ing loads. 

The  frictional  resistance  per  square  inch  is  the  product  of  the  coefficient  of 
friction  times  the  load  per  square  inch  on  the  projected  area  of  the  bearing. 
Thus,  if  this  product  is  a  constant  one  factor  must  vary  inversely  as  the  other; 


SLIDING  FRICTION 


21 


a  high  load  will  give  a  low  coefficient  and  a  low  load  a  high  coefficient.     This  is 
true  for  well  lubricated  surfaces. 

For  ordinary  lubrication  the  coefficient  is  more  nearly  constant  under  vary- 
ing loads;  the  friction  resistance  then  varies  directly  as  the  load.  This  is  shown 
by  Table  8  taken  from  the  results  of  the  experimental  work  of  Beauchamp 
Tower,  Proc.  Inst.  M.  E.,  1883. 


Nominal 

load 

lb. 

per  sq.  in. 

Actual 

load 

lb. 

per  sq.  in. 

Temper- 
ature 
Fahr. 

100  rev. 

I  OS    ft. 

per  min. 

lb. 

ISO  rev. 

IS7  ft. 

per  min. 

lb. 

200  rev. 

209  ft. 

per  min. 

lb. 

250  rev. 
262  ft. 
per  min. 
lb. 

300  rev. 
314  ft. 
per  min. 
lb. 

3 so  rev. 

366  ft. 

per  min. 

lb. 

400  rev. 

419  ft. 

per  min. 

lb. 

328 
310 
293 
275 
258 
20S 
153 
100 

582 
551 

S20 
498 

4S8 
364 
272 
178 

82° 
76° 

77° 
78° 
82° 
74° 
75° 

3-5 

3.06 

3.06 

2.49 

2.44 

1.78 

1-473 

1.093 

3-35 

2.84 

2.84 

2.62 

2.28 

I.73S 

I.S6 

I. 225 

3-21 

3  06 
2.84 
2.84 
2.17 
1. 60s 
1. 60s 

..33 

3.06 

3.c6 

2.49 

2.89 

2.X4S 

X.37 

0.992 

. 

2.1 
1.75 
1. 81 
1.44 

2.13 
2.04 
1.89 
1. 54 

1.56 
1 .0 

The  chord  of  the  arc  of  contact  of  the  brass  =  2  1/4  in. 

The  nominal  load  per  square  inch  is  the  total  load  divided  by  4X6. 

The  actual  load  per  square  inch  is  the  total  load  divided  by  2  1/4X6. 

The  results  with  the  actual  load  of  582  lb.  per  square  inch  were  obtained  with 
difficulty,  and  the  bearing  seized  with  that  load  after  running  for  a  short  time. 

The  lubricating  pad  consisted  of  a  piece  of  felt  pressing  against  the  journal,  and  resting 
on  worsted  immersed  in  a  tin  box  full  of  oil. 


Table  8.- 


-Relationship  of  Load  and  Frictional  Resistance  for  Ordinary 
Lubrication 


Mr.  Stroudley  with  pad  lubrication  on-  railroad  car  axles  on  the  Brighton 
Railway  in  England  obtained  the  result  of  Table  9  which  still  further  supports 
the  law. 


Load  per  square  inch  on 
brass  in  pounds 


Coefficient  of  friction 


293 
251 
237 


0.00792 
o . 0080 
0.0077 
o . 007  2 


Friction  resistance  per  square 
inch  in  pounds 


2.639 

2.34 
1-93 
1.70 


Table  9. — Relationship  of  Load  and  Frictional  Resistance  as  Determined  by 

Stroudley. 


22 


BEARINGS  AND  THEIR  LUBRICATION 


DISCUSSION   OF   THE   FOURTH   LAW 

Turning  to  the  fourth  law,  A.  M.  Wellington,  Trans.  A.S.C.  E.,  Vol.  XIII, 
found  from  experiment  on  journals  revolving  at  very  low  velocities  that  the 
friction  under  those  conditions  was  very  great  and  nearly  constant  under  widely 
varying  conditions  of  lubrication,  load  and  temperature.  As  the  speed  increased 
the  friction  fell  slowly  and  regularly,  returning  again  to  the  original  amount 
when  the  velocity  was  reduced  to  the  same  rate.  The  tabulation  below  shows 
some  of  the  results. 

Feet  per  minute,  2.16     3.33        4.86  8.82        21.42     35-37      53.oi      89.28         106.02 

Coefficient  of  friction,  0.118  0.094    0.0705     0.0685     0.055     0.047      0.04      0.035     0.3   0.0255 

This  shows  a  regular  decrease  in  the  value  of  the  coefficient  of  friction  with 
increase  of  speed. 

The  results  of  experiments  made  by  Professor  Kimball,  see  page  1904 
Americin  Journal  of  Science,  March,  1878,  show  this  same  law  of  the  decrease 
of  the  coefficient  of  friction  with  the  increase  of  speed.  The  follov^ng  gives 
some  of  Professor  Kimball's  results: 

Velocity  in  ft.        i  3  5  7  10  15  20  30  40  60  80  100 

per  min., 
Coefficient,  0.15      0.122     0.114     0.093    0.079  0.066  0.058  0.054  0.053   0.052  0.051     0.05 


These  experiments  were  made  with  journal  bearings  and  show  that  the 
friction  was  reduced  67  per  cent,  with  an  increase  of  velocity  from  i  to  100  ft. 
per  minute.  After  this  limit  of  velocity  100  ft.  per  minute  has  been  passed,  the 
coefficient  of  friction  varies  approximately  with  the  square  root  of  the  velocity. 


Origin  of  data 


Observed. . 
Calculated. 


Observed.. 
Calculated. 


Load 

lb. 

per  sq.  in. 


Velocity  in  feet  per  minute 


209 


262 


314 


366 


419 


520 


468 


0.0013      0.0014        0.0015        0.0017 
0.00145      0.00159      0.00172 


0.0015 
0.00184 


471 


o.ooi    0.0012  [  0.0013    0.0014   0.0015    0.0017 
I  0.0II8   0.00123   0.00132   0.00I4I   0.0015 


0.002 
0.00195 


Observed.. 
Calculated. 


415 


0.0014  !  0.0015      0.0017      0.0019      0.0021      0.0024 

0.00157       0.00172   !   0.00188   I   0.00198   !   0.0021 


Table  10. — Comparison  of  Observed  and  Calculated  Coefficients  of  Friction  for 
Velocities  over  100  Ft.  per  Minute. 


SLIDING  FRICTION 


23 


Table  10  shows  us  this  fact.  The  observed  coefficients  of  friction  for  various 
velocities  are  taken  from  some  of  Mr.  Tower's  experiments.  The  comparable 
calculated  coefficients  were  computed  from  the  ratio  of  the  square  roots  of  the 
velocities.     The  agreement  is  very  close. 

For  the  experimental  work  to  determine  the  coefficient  of  friction  at  high 
speeds,  we  are  indebted  to  Thurston,  Stribeck  and  Lasche.  In  Traction  and 
Transmission,  January,  1903,  Lasche  thus  sums  up  the  results  of  these  experi- 
ments: From  500  ft.  per  minute  to  800  ft.  per  minute  the  experiments  of  Stri- 
beck show  that  the  rise  in  the  value  of  the  coefficient  of  friction  is  slow,  and 
approaches  the  result  of  the  Thurston  experiment  expressed  in  the  formula 


~ 

~ 

— 

-w 

Ste 
hit< 

3land 

1  IV"*' 

,1 

N 

ckelS 

teelar 

d 

lit! 

Utl 

Nickel  Steel  and 

^ 

Mild  Ste 

el  a 

nd 

1  0.020 

V 

in\ 

e  ^ 

let 

al. 

^JjkelSt 

eel 

0 

aie 

•5  0.010 

1 



jL 

^ 

^ 

^ 

^ 

^ 

;=:: 

zz 

= 

= 

^^ 

== 

-=4 

— 1 







0 

n 

— 

-^ 

^^'. 

= 

— 

= 

— 

1 

r 

= 

— 

— 

— 

n 

~ 

— ■ 

— 

1000  2000  3000  4000 

Circumferential  Journal  Speed  in  Feet  per  Minute. 


5000 


FIG.    2. 


CURVES    FROM  LASCHE'S  EXPERIMENTS  SHOWING  THE  VARIATION  IN  THE  COEFFICIENT 
OF   JOURNAL    FRICTION   WITH   VELOCITY. 


f=\/V.  At  a  still  higher  speed  the  influence  of  the  velocity  disappears  and  at 
1600  ft.  per  minute  the  curve  of  the  values  of  the  coefficient  is  only  slightly 
higher  than  at  800  ft.  per  minute.  The  experiments  of  Lasche  further  give  the 
conclusion  that  for  velocities  of  2000  ft.  per  minute  and  over  the  coefficient  of 
friction  is  practically  independent  of  the  velocity. 

There  is  a  break  in  these  results  between  a  speed  of  1600  ft.  per  minute  and 
2000  ft.  per  minute,  but  from  a  consideration  of  the  plotted  curve  of  Lasche's 
work  we  seem  justified  in  assuming  that  for  velocities  of  1600  ft.  per  minute  and 
over  the  coefficient  of  friction  is  practically  independent  of  the  velocity. 

Fig.  2  shows  the  general  shape  of  the  coefficient  of  friction  curves  plotted 
against  velocities  of  rubbing  as  abscisses. 


DISCUSSION  OF   THE  FIFTH  LAW 

Law  5  states  that  the  variation  of  friction  with  temperature  is  approximately 
in  inverse  ratio.  The  following  taken  from  Mr.  Tower's  experiments  at  a 
velocity  of  262  ft.  per  minute  with  an  oil  bath  show  this  law: 


24 


BEARINGS  AND  THEIR  LUBRICATION 


Temperature  in  Fahrenheit 
degrees 


Observed  coefficient  of 
friction 


Calculated  coefhcient  of 
friction 


no 

lOO 

90 
80 
70 
60 


o . 0044 
0.0051 
0.006 

0.0073 

0.0092 
O.OII9 


0.00451 
0.00518 
o . 00608 

0.00733 

0.00964 
0.01252 


However,  this  law  does  not  hold  good  for  imperfect  lubrication,  that  is, 
for  pad  or  syphon  lubrication.  Then  the  coefficient  of  friction  diminishes  more 
rapidly  for  a  given  temperature  according  to  a  gradually  decreasing  scale,  until 
a  normal  temperature  is  reached.  This  is  shown  by  the  result  of  Mr.  Stroud- 
ley's  experiments  with  a  pad  of  rape-seed  oil. 


Temperature  Fahrenheit 
degrees 

Coefficient  of  friction 

Decrease  of  coeifjcient 

10;                       

0.022 

0.018 

0.016 

0.014 

0.0125 

0.0115 

O.OII 

0.0106 
0.0102 

0.004 
0.002 

no 

lie 

0.002 

120                 

0.0015 

O.OOI 

i2t: 

130 

i^e 

0.0005 
0 . 0004 

0   0002 

140.          . .         

141; 

THICKNESS  OF  THE  OH.  FH^M 

A  series  of  experiments  made  by  Goodman  to  determine  the  thickness  of 
the  oil  film  showed  that  with  pad  lubrication  on  first  starting  the  journal,  the 
thickness  of  the  film  was  0.0013  i^-  As  the  speed  increased  and  the  film  became 
better  developed  its  thickness  increased  to  0.0029  in. 

In  the  American  Machinist  for  September  17,  1903,  page  13 16,  Herbert  S. 
Moore,  writing  in  regard  to  a  film  formed  with  a  heavy  engine  oil,  states: 

''It  was  at  first  attempted  to  get  a  perfect  film  in  the  Olsen- Carpenter  fric- 
tion machine,  in  which  the  test  bearing  was  a  snug  fit  all  around  the  journal. 
It  was  found  impossible  to  get  such  a  film  except  at  rare  intervals  of  short  dura- 
tion, whereas  with  the  altered  Thurston  machine,  in  which  the  diameter  of  the 


SLIDING  FRICTION  25 

bearing  was  0.003  i^-  greater  than  that  of  the  journal,  this  film  was  always 
readily  formed  and  preserved." 

If  the  film  was  uniform  its  thickness  would  thus  be  0.0015  in. 

BREAKING-DOWN  POINT  OF  THE  OIL  FILM 

The  difficulty  of  lubricating  bearings  under  high  pressures  is  well  known. 
The  reason  is  evident.  As  the  pressures  are  increased  the  film  of  oil  is  estab- 
lished and  maintained  with  increasing  difficulty.  If  the  pressures  are  increased 
to  a  sufl&cient  amount  the  film  of  oil  finally  breaks  down,  and  the  journal  and 
bearings  come  into  metallic  contact. 

Herbert  S.  Moore,  on  page  1283,  American  Machinist  for  September  10, 
1903,  gives  a  formula  based  on  experimental  investigations,  to  determine  the 
pressure  at  which  a  film  of  machinery  oil  will  break  down.     This  formula  is 

in  which  p  is  the  intensity  of  the  load  in  pounds  per  square  inch,  and  v  is  the 
rubbing  velocity  of  the  journal  in  feet  per  minute. 

The  experimental  work  consisted  in  connecting  into  an  electric  circuit  the 
rotating  journal  and  its  bearing.  As  long  as  the  oil  film  was  properly  main- 
tained the  resistance  of  this  circuit  was  very  high.  As  soon,  however,  as  the  oil 
film  broke  down,  giving  metallic  contact  between  journal  and  bearing,  the 
resistance  of  the  circuit  fell  to  almost  nothing. 

PRESSURE  OF  THE  OIL  FILM 

If  a  journal  is  completely  fluid-borne  it  follows  that  the  oil  film  must  be 
under  a  considerable  pressure  in  order  to  support  the  weight  of  the  journal  and 
its  attached  parts.  That  such  is  a  fact  was  demonstrated  by  an  experiment  by 
Mr.  Towers  reported  in  the  Proc.  Inst.  M.  E.  for  1885,  P^g^  5^-  I^  these  ex- 
periments a  half  bearing  or  brass  was  used,  having  a  length  of  6  in.  and  a 
chord  length  of  the  arc  of  contact  of  3.9  in.  Three  longitudinal  holes  were 
drilled  a  little  more  than  half  way  of  the  length  of  the  bearing,  and  spaced  as 
indicated  by  the  letters  a,  d  and  g  in  Fig.  3. 

At  9  points,  as  indicated  by  the  letters  in  the  figure,  holes  were  drilled 
from  the  inside  of  the  bearing  into  these  three  longitudinal  channels.  Each 
channel  was  connected  with  a  pressure  gage,  and  one  hole  at  a  time  was  left 
open  from  the  channel  into  the  inside  of  the  bearing.  By  this  means  pressure 
readings  were  taken  at  nine  different  points,  three  on  the  side  of  the  bearing 
where  the  journal  entered,  three  on  the  middle  line,  and  three  on  the  side  of  the 
bearing  where  the  journal  left.  Furthermore,  as  six  of  these  holes  were  between 
a  middle  transverse  plane  and  the  end,  and  since  it  is  fair  to  assume  that  the 


26 


BEARINGS  AND  THEIR  LUBRICATION 


same  pressure  would  exist  on  the  other  side  of  this  middle  plane,  it  can  be  said 
that  the  experiments  covered  15  points  on  the  surface  of  the  bearing. 

The  lubrication  was  by  means  of  an  oil  bath,  the  journal  being  about  half 
submerged.  The  speed  of  rotation  was  kept  uniform  at  150  revolutions  per 
minute,  and  the  temperature  was  held  constant  at  90°  F.  The  total  load  on 
the  bearing  was  8008  lb.,  and  the  observed  oil  pressure  at  the  9  points 
investigated  were  as  shown  by  the  following: 

For  longitudinal  plane  at  entering  side  of  the  journal  a  310,  h  335,  c  370; 
for  middle  longitudinal  plane  d  565,  e  615,/  625;  for  longitudinal  plane  at 
leaving  side  of  the  journal  g  430,  h  480,  i  500. 

A  summation  of  the  total  pressure  on  the  bearing  from  these  observations 
gives  7988  lb.  This  compares  very  closely  with  the  actual  pressure,  which 
was  8008  lb.,  the  small  difference  being  probably  due  to  inaccuracies  in 
observations. 


c»i       i       i       ■ 

, 

III' 

s  — 

E 

1         1         1 
1        1        1 

[i       \h       [p" 
1 ]—-■—■ 

1                  1 
1         1         1 

.1 

1         1         1 

FIG.    3      DIAGRAM  SHOWING  POINTS  AT  WHICH  OIL  PRESSURES  WERE  TAKEN. 


These  results  present  several  interesting  facts.  The  length  of  the  bearing 
was  6  in.,  and  the  length  of  its  chord  of  contact  3.9  in.,  giving  a  projected  area 
of  23.4  sq.  in.  The  applied  load  was  8008  lb.,  which  divided  by  the  pro- 
jected area  gives  an  intensity  of  pressure  per  square  inch  of  about  311  lb. 
The  highest  observed  pressure  at  the  point/,  or  center  of  the  bearing  area,  was 
625  lb.,  or  double  the  computed  intensity  of  pressure. 

The  observed  pressures  at  the  points  g,  h  and  i  on  the  leaving  side  of  the  bear- 
ing are  considerably  greater  than  the  corresponding  pressures  at  the  points  a,  h 
and  c  on  the  entering  side.  This  shows  the  shifting  of  the  point  of  maximum 
pressure  in  the  bearing  from  a  point  diametrically  below  or  above  the  axis  of 
rotation,  to  a  point  with  the  direction  of  rotation.  See  the  following  section 
devoted  to  a  discussion  of  the  points  of  maximum  and  minimum  pressure  in  an 
oil  film. 


SLIDING  FRICTION  27 

The  existence  of  these  pressures  in  an  oil  film  emphasizes  the  necessity  of 
using  good  care  and  judgment  in  selecting  the  points  where  lubricant  is  to  be 
introduced.  Obviously  oil  cannot  be  introduced  into  a  bearing  unless  it  is 
under  a  greater  pressure  than  the  pressure  in  the  oil  film  at  the  point  of  entry. 
Some  failures  in  elaborate  oiling  systems  can  be  traced  to  this  very  fact;  the 
attempt  to  feed  oil  at  a  point  in  the  bearing  where  the  oil  film  was  under  a 
greater  pressure  than  the  head  of  the  feed. 

Referring  again  to  Mr.  Tower's  experiment,  he  found  that  by  varying  the 
load  the  oil  pressures  rose  and  fell  in  exact  proportion.  Putting  on  weight 
increased  the  oil  pressures,  and  removing  weight  decreased  them.  To  deter- 
mine the  effect  of  speed  on  this  oil  pressure  the  journal  speed  was  reduced  from 
150  revolutions  per  minute  to  20,  or  as  the  diameter  of  the  shaft  was  4  in.,  from 
about  150  ft.  per  minute  to  about  20  ft.  per  minute.  The  gage  pressures  of  the 
oil  film  were  exactly  the  same  at  the  lower  speeds  as  at  the  higher.  Reynolds* 
theory  and  Kingsbury's  experiments  show,  however,  that  the  distribution  of  the 
pressure  does  vary  with  speed. 

As  there  are  points  of  maximum  pressure  in  the  oil  film  in  a  bearing,  there 
are  likewise  points  of  minimum  pressure,  and  these  may  be  below  the  pressure 
of  the  atmosphere,  or  at  a  slight  vacuum.  This  is  true  for  arcs  of  contact 
from  360  to  less  than  180  degrees.  At  page  1282  of  the  American  Machinist, 
September  30,  1903,  Herbert  F.  Moore  describes  an  experiment  with  a  small 
dynamo  bearing,  in  which  the  bearing  sucked  oil  against  a  head  of  6  in.  We 
quote:  *'By  simply  leading  a  pipe  from  a  reservoir  of  oil  to  that  part  of  the 
bearing  where  the  pressure  would  be  negative,  if  anywhere,  it  is  found  that  the 
bearing  would  suck  in  oil  from  a  reservoir  6  in.  below  and  thus  lubricate  itself." 

POINTS  OF  MAXIMUM  AND  MINIMUM  PRESSURE  AND 
NEAREST  APPROACH 

In  the  earlier  discussions  of  the  relative  positioning  of  a  journal  and  its 
mating  bearing  when  the  journal  was  in  motion  with  ample  lubrication  present, 
it  was  stated  that  the  journal  rolled  back  so  that  the  point  of  nearest  approach 
was  on  the  on  side  of  the  bearing  and  at  an  angular  distance  from  the  vertical — 
assuming  the  load  to  act  vertically  with  the  bearing  supporting  the  journal — 
equal  to  the  angle  of  repose  of  the  system.  This  theory  has  been  shown 
to  be  fallacious  by  the  experimental  work  of  Tower  and  analytical  work  of 
Reynolds. 

In  an  ordinary  journal  bearing  where  there  is  some  looseness,  when  the 
journal  is  at  rest  it  lies  at  the  bottom  of  the  bearing  and  the  point  of  nearest  ap- 
proach, in  this  case  contact,  lies  in  a  vertical  passing  through  the  centers  of  both. 
The  rolling  action  outlined  in  the  paragraph  above  may  take  place  during  the 


28 


BEARINGS  AND  THEIR  LUBRICATION 


190 

180 

170 

160 

.150 

|140 

Kl30 

S120 


o 

^ 

, — 



- — 

— 

— 

' 

^^• 

pr^ 

^^r 

J/ 

k 

V 

/ 

/ 

/ 

/ 

/ 

80 
70 
60 
50 
40 
30 

^l< 

k 

pr-1 

^>. 

*^ 

V. 

r-- 

*^> 

ki 

'^c^,V 

r^J 

r"-^r 

N: 

-^. 

■ — .^ 

Th 

nnes 

tPo 

1      1 

int  of  Film 

^ 

Oil 

"■^ 

' 

^^ 

=^ 

^ 

::^ 

^^ 

N 

CLc 

^ 

;^ 

^ 

-. . 

\ 

-.* 

^fnt 

"'^- 

_^ 

3 

20 
10 

p^ 

tax. 

Pres 

surp 

o^- 

~1 





— 







> 

n 

• 

0.0007 


0.0006 


0.0005 


0.0004 


0.0003 


0.0002 


0.0001 


800 


1200 


IGOO  2000         2400         28U0         3;i00 

Revolutions  per  Minute 


3600    4000 


4400 


400  800         1200         1600         2000        2400        2800         3200         3600        4000        4400 

Revolutions  per  Minute 

FIG.    4.      RELATION   BETWEEN  POINTS   OF   MINIMUM  AND   MAXIMUM   PRESSURE 
AND    POINT   OF   NEAREST  APPROACH. 


SLIDING  FRICTION  29 

period  of  starting,  but  if  there  is  any  appreciable  amount  of  lubricant  present, 
as  soon  as  sliding  takes  place  the  point  of  nearest  approach  and  the  point  of 
maximum  pressure  (which  are  not  identical)  are  thrown  to  the  off  side  of  the 
vertical  through  the  centers  of  journal  and  bearing.  See  the  upper  diagram  in 
Fig.  4. 

The  angular  amount  of  this  shifting  is  dependent  on  the  fit  and  finish  of  the 
surfaces,  the  quality  of  lubrication,  the  speed,  and  the  pressure.  With  a  vari- 
able load  and  constant  speed,  as  the  load  increases  the  point  of  nearest  approach 
swings  upward  continuing  until  it  reaches  a  position  nearly,  but  not  quite,  at  the 
extremity  of  the  horizontal  axis.  With  a  further  increase  of  load,  it  swings  back 
to  a  position  found  by  Reynolds  to  be  about  40  degrees  from  the  vertical.  With 
a  still  further  increase  of  load,  the  oil  film  is  ruptured  and  the  conditions  are 
changed  from  that  of  perfect  lubrication  to  that  of  imperfect  lubrication. 

Quoting  from  Reynolds'  paper,  Phil.  Trans.  (Royal  Society,  London)  1886: 

''The  circumstances  which  determine  the  greatest  load  which  a  bearing  will 
carry  with  complete  lubrication,  that  is,  with  a  film  of  oil  extending  between 
brass  and  journal  throughout  the  entire  area,  are  definitely  shown  in  the  theory. 

"  The  effect  of  increasing  the  load  beyond  a  certain  small  value  being  to  cause 
the  brass  to  approach  nearer  to  the  journal  at  a  point  on  the  off  side  which  moves 
toward  the  vertical  as  the  load  increases  and  when  the  load  is  such  that  the  least 
separating  distance  is  about  half  the  difference  of  the  radii,  the  angular  position 
of  the  point  of  nearest  approach  is  40  degrees  to  the  off  side  of  the  vertical 
through  the  middle  of  the  brass.  At  this  point  the  pressure  in  the  oil  film  is 
everywhere  greater  than  at  the  extremities  of  the  brass,  but  when  the  load 
further  increases  the  pressure  toward  the  extremity  on  the  off  side  becomes 
smaller  or  negative.  This,  when  suflScient,  will  cause  rupture  in  the  oil  film, 
which  will  then  only  extend  between  the  brass  and  journal  over  a  portion  of  the 
whole  area  and  a  smaller  portion  as  the  load  increases.  Thus  since  the  amount 
of  negative  pressure  which  the  oil  will  bear  depends  on  circumstances  which 
are  uncertain,  the  limit  of  the  safe  load  for  complete  lubrication  is  that  which 
causes  the  least  separating  distance  to  be  half  the  difference  of  the  radii  of  brass 
and  journal." 

An  examination  of  Reynolds'  equations  shows  that  the  load  and  speed  are 
inversely  related;  that  is,  an  increase  of  load  has  an  effect  upon  the  form  of  the 
film  of  the  same  kind  as  a  decrease  of  speed,  and  vice  versa. 

That  practice  agrees  with  this  theory  is  shown  by  a  phenomenon  observed 
by  Tower  and  by  the  position  of  the  wear  in  car  journal  brasses.  Tower  men- 
tions that  the  journal  having  been  run  in  one  direction  until  the  initial  tendency 
to  heat  had  entirely  disappeared,  on  being  reversed  immediately  began  to  heat 
again;  but  this  effect  stopped  when  the  process  had  been  often  repeated.  The 
fact  being  that  running  in  one  direction,  the  brass  had  been  worn  to  the  journal 


30  BEARINGS  AND  THEIR  LUBRICATION 

only  on  the  off  side  for  that  direction,  so  that  when  the  motion  was  reversed  the 
new  off  side  acted  like  a  new  brass. 

In  car  brasses  the  point  of  wear  is  always  forward  or  on  the  off  side. 

The  relative  relations  of  the  points  of  maximum  pressure,  minimum  pres- 
sure, and  nearest  approach,  are  shown  by  the  curves  of  Fig.  4.  These  are 
taken  from  a  paper  of  Professor  Kingsbury,  published  in  the  Journal  of  the 
American  Society  of  Naval  Engineers,  Vol.  IX,  No.  2,  1897,  and  apply  to  an 
experiment  with  an  air-lubricated  bearing,  running  under  a  constant  load  and  at 
variable  speed. 


SECTION  II 

COEFFICIENTS  OF  FRICTION  OF  JOURNAL,  COLLAR,  STEP  AND 

GUIDE  BEARINGS 

From  the  discussion  of  imperfect  lubrication,  that  is  the  ordinary  lubrication 
to  machinery  journal  bearings,  it  is  evident  that  no  fixed  values  can  be  given  for 
the  coefficient  of  friction.  Any  values  that  are  given  must  be  for  definite  lubri- 
cants, and  not  for  surfaces  in  contact.  With  perfect  lubrication  and  moderate 
velocity  the  coefficient  may  be  as  low  as  o.ooi.  With  very  low  velocities  and 
high  pressures  the  coefficient  approaches  that  for  greasy  or  unlubricated  sur- 
faces, a  minimum  value  for  which  may  be  taken  as  0.15.  Tests  of  new  well- 
fitted  bearings  have  given  values  of  0.30  and  more.  These  values  are  the  ones  to 
use  in  estimating  the  starting  moment  for  heavy  machinery. 

On  page  105  of  Elements  of  Machine  Design,  by  Kimball  and  Barr,  the  coeffi- 
cients of  friction  to  be  used  in  the  design  of  ordinary  machinery  are  given  as 
follows:  A  fair  average  range  for  pressures  from  50  to  500  lb.  and  velocities 
from  50  to  500  ft.  per  minute  is  from  0.02  to  0.008,  and  for  purposes  of  design  of 
ordinary  machinery,  may  be  taken  at  0.015. 

Perhaps  the  lowest  coefficient  of  friction  that  has  been  reported  from  authen- 
tic tests  will  be  found  on  page  149,  Trans.  A.S.  M.  E.,  Vol.  LXXIV.  There 
Professor  Kingsbury  presents  a  table  giving  coefficients  of  friction  for  sperm  oil 
under  a  pressure  of  340  lb.  per  square  inch  at  a  temperature  of  90°  F.,  and 
varying  rubbing  speeds. 


42  feet  per  minute 

64  feet  per  minute 

loi  feel,  per  minute 

0 . 000906 

0.000915 

0 . 0008S6 

0.000915 

0.001171 
0.001158 
0.001158 
O.OOI 1 25 

0.00146 
0.00155 
0.00151 

On  page  146  of  the  same  paper  a  value  of  0.00053  is  given. 
The  lowest  coefficient  of  which  the  author  has  learned  is  0.00046.     This  was 
also  obtained  by  Professor  Kingsbury  in  a  test,  the  results  of  which  have  never 

31 


32 


BEARINGS  AND  THEIR  LUBRICATION 


before  been  published.  The  lubricant  was  spindle  oil,  pressure  340  lb.  per 
square  inch,  temperature  135°  F.,  and  a  rubbing  speed  of  42  4  ft.  per  minute. 
To  give  a  definite  idea  of  the  value,  the  coefficients  of  friction  for  general 
bearings  and  the  way  in  which  they  vary  under  different  conditions.  Tables  11  to 
19,  inclusive,  are  presented.  They  are  all  from  the  experiments  of  Mr.  Tower, 
made  under  the  auspices  of  the  Institution  of  Mechanical  Engineers.     In  using 

0.03 


.H  0.02 

u 


I  0.01 


, 

•- 

_ 

Imperial  0. 

Oil 

«k 

■^ 

0 

o- 

RaieS 
Sperm 

3ed  Oil. 
Dil. 

oil 

^ 

^^ 

- 

— G— 

_ 

-v>-^ 

^ 

-o— 

— ♦« 

68 


101  140 

Temperature  of  Bearing  t  in  Fahrenheit  Degrees. 


176 


212 


Conditions:  Pressure  on  bearing,  Q2  pounds  per  square  inch;  rubbing  speed,  195  feet 
per  minute;  Bearing  flooded  with  oil,  about  0.8  quart  per  minute;  journal  of  nickel  steel, 
bushes  of  white  metal,  diameter  about  10  inches,  length  about  4  1/2  inches. 

riG.    5.       COTZFFICIENTS    OF    FRICTION   FOR   DIFFERENT    LUBRICANTS    (Laschc). 
0.05 


0.04 


g  0.03 

o 

S  0.02 


..... 

i 

\\ 

■\ 

■'\ 

V 

\S 

• 

Imperia 
Rape  S€ 

1  0.  Oil 
'ed  Oil 

^ 

s 

Spe 

rm( 

3il 

^^ 

*^ 

=^ 

^=7:^ 

e=rr 

■    — ■ 

■■ 

_..J| 

0.01 


0  71  142  213 

Pounds  per  Square  Inch. 
Conditions:     Rubbing  speed,  IQ5  feet  per  minute;  temperature,  122  degrees  Fahrenheit; 
bearing  flooded  with  oil,  about  0.8  quart  per  minute;  journal  nickel  steel,  bushings  white 
metal,  diameter  about  10  inches,  length  about  4  1/2  inches. 

FIG.  6.     coefficients  of  friction  for  different  lubricants  (Lasche). 

any  values  from  these  tables  it  must  be  clearly  recognized  for  what  kind  and 
degree  of  lubrication  they  are  applicable. 

Figs.  5  and  6  are  reproduced  from  Traction  and  Transmission,  January, 
1903,  and  show  results  of  Lasche's  experiments.  The  "imperial  oil"  there 
mentioned  is  a  Russian  mineral  oil  of  good  quality  for  lubricating  purposes. 


COEFFICIENTS  OF  FRICTION 


33 


Nominal 

load 

lb. 
per  sq.  in. 

Speeds 

100  rev. 

105  ft 

per  min. 

150  rev.   200  rev. 

157  ft.    20;  ft. 

per  min.  per  min. 

i 

250  rev. 

262  ft. 

per  min. 

300  rev. 

314  ft. 

per  min. 

350  rev. 

366  ft. 

per  min. 

400  rev. 

419  ft. 
per  rain. 

Lb. 

625 

520 

415 
310 
205 
100 

0.0013  1  0.00139  0.00147  0.00157 
0.00123  0.00139  0.0015    0.00161 

0.00123    O.OOIA-2    0.0016    '  0.00176 

0.00165 
0.0017 

0.00178 

0.0019    0.002 

0.00178 
0.00334 

0.00142 
0.00205 
0.00415 

0.0016 
0.00235 
0 . 00494 

0.00184 
0.00269 
0.00557 

0.00207 
0.00298 
0.0062 

D. 00225 
0.00328 
0.00676 

0.00241 

0.0035 

0.0073 

Conditions:  A  4-in,  journal  of  steel  6  in.  long  fitted  on  its  upper  side  with  a  gun 
metal  brass.,  or  half  box,  embracing  somewhat  less  than  one-half  the  journal  circumfer- 
ence— the  chord  of  the  arc  contact  equal  to  3.92  in. — and  when  running  kept  at  a  constant 
temperature  of  90°  F. 


Table  ii. — Coefficients  of  Friction  for  A  Bath  of  Mineral  Oil- 
Constant     {Tower). 


-Temperature 


Nominal 

load 

lb. 

per  sq.  in. 

Speeds 

100  rev. 

150  rev. 

200  rev. 

250  rev. 

300  rev. 

350  rev. 

400  rev. 

450  rev. 

105  ft. 
per  min. 

157  ft. 
per  min,. 

209  ft. 
per  min. 

262  ft. 
per  min. 

314  ft. 
per  min. 

366  ft. 
per  min. 

419  ft. 
per  min. 

471  ft. 
per  min. 

Lb. 

625 

O.OOI 

0.0012 

0.0014 

0.0014 

0.0016 

0.0018 

0.002 

520 

0.0014 

0.0016 

0.0018 

0.0019 

0.002 

0.0021 

0.0022 

415 

0.0016 

0.0019 

0.0021 

0.0023 

0.0025 

0.0026 

0.0027 

310 

0.002 

0.0022 

0.0026 

0.0029 

0.0032 

0.0035 

0.00^8 

0.004 

205 

0 . 0026 

0.0034 

0 . 0040 

0.0047 

0.0053 

0.0058 

0.0062 

0.0066 

153 

0.0028 

0.0038 

0 . 0048 

0.0057 

0.0065 

0.0071 

0.0077 

0.0083 

100 

0  0054 

0.0076 

0 . 0094 

0.0109 

0.0123 

0.0133 

0.0142 

0.0151 

Conditions:  A  4-in.  journal  of  steel  6  in.  long  fitted  on  its  upper  side  with  a  gun 
metal  brass,  or  half  box,  embracing  somewhat  less  than  one-half  the  journal  circumfer- 
ence— the  chord  of  the  arc  contact  equal  to  3.92  in. — and  when  running  kept  at  a  constant 
temperature  of  90°  F. 


Table  12. — Coefficient  of  friction  for  a  Bath  of  Mineral  Grease- 
Constant  {Tower). 
3 


-Temperature 


34 


BEARINGS  AND  THEIR  LUBRICATION 


Nominal 
load 
lb. 

per  sq.  in. 

Speeds 

100  rev. 

105  ft. 

per  min. 

150  rev. 

157  ft. 

per  min. 

200  rev. 

2og  ft. 

per  min. 

250  rev. 
262  ft. 
per  min. 

300  rev. 
314  ft. 
per  min. 

350  rev. 
366  ft. 
per  min. 

400  rev. 

419  ft- 
per  min. 

450  rev. 
471  ft. 
per  min. 

,Lb. 
'S20 

415 
310 
205 

153 
100 

Seized 

! 

0.0015 

O.OOII 

0.0016 
o.ooig 
0.003 

0.0017 
0.0012 
0.0018 
.0.0023 
0 . 0038 

0.0018      O.ooig      0.002 

0.0014        0.0016        O.OOI7 

0.0021 
0.0018 
0.0025 
0.0035 
0.0061 

0.0021 
o.ooig 
0.0027 
0.0037 
0 . 0064 

0.0013 
0.0016 
0.0025 

0.0021 
0.0028 
0.0044 

0.0023 
0.0030 
0.0051 

0.0024 
0.0033 
0.0057 

Conditions:  A  4-in.  journal  of  steel  6  in.  long  fitted  on  its  upper  side  w^ith  a  gun 
metal  brass,  or  half  box,  embracing  somewhat  less  than  one-half  the  journal  circumfer- 
ence— the  chord  of  the  arc  contact  equal  to  3.92  in. — and  when  running  kept  at  a  constant 
temperature  go°  F. 

Table  13. — Coefficient  of  Friction  for  a  Bath  of  Sperm  Oil — Temperature  Con- 
stant {Tower). 


N^nm  1  n  a  1 

Speeds 

load 

lb. 

per  sq.  in. 

100  rev. 

105  ft. 

per  min. 

150  rev. 

157  ft. 

per  min. 

200  rev. 

2og  ft. 

per  min. 

250  rev. 
262  ft. 
per  min. 

300  rev. 
314  ft. 
per  min. 

350  rev. 
366  ft. 
per  min. 

400  rev. 
4ig  ft. 
per  min. 

4^0  rev. 
471ft. 
per  min. 

Lb. 

415 

•  310 

205 

153 

100 

0 . ooog 
0 . 001 2 
0  0014 
0.0020 
0.0027 
0.0042 

O.OOI 

0.0014 
0.0017 
0.0023 
0.0032 
0.005 

O.OOII 

0.0015 
0.002 
0 .0028 
0.0037 
0.006 

0.0013 
0.0016 
0.0022 
0 . 003 I 
0 . 004 I 
0 . 0067 

0.0015 

0.0018 

0.0025 

0.0034 

0.005 

0.0076 

0.0015 
o.ooig 
0 . 0026 
o.oo3g 
0.0051 
0.0081 

0.0017 

0.0029 
0.0042 
0.0052 
0.009 

0.0017 
0.0022 
0.003s 

Conditions:  A  4-in.  journal  of  steel  6  in.  long  fitted  on  its  upper  side  with  a  gun 
metal  brass,  or  half  box,  embracing  somewhat  less  than  one-half  the  journal  circumfer- 
ence— the  chord  of  the  arc  contact  equal  to  3.92  in. — and  when  running  kept  at  a  constant 
temperature  of  90°  F. 


Table  14. — Coefficients  of  Friction  for  Lard  Oil  Bath — Temperature  Constant 

{Tower). 


COEFFICIENTS  OF  FRICTION 


35 


Speeds 


Temper- 

ature 

100  rev. 

150  rev. 

200  rev. 

250  rev. 

300  rev. 

350  rev. 

400  rev. 

450  rev. 

Fahr. 

105  ft. 

157  ft. 

209  ft. 

262  ft. 

314  ft. 

366  ft. 

419  ft. 

471  ft. 

per  min. 

per  mm. 

per  mm. 

per  mm. 

per  mm. 

per  mm 

per  mm. 

per  mm. 

120° 

0.0024 

0.0029 

0.0035 

0.004 

0 . 0044 

0.0047 

0 . 005 I 

0.0054 

110° 

0.0026 

0.0032 

0.0039 

0.0044 

0.005 

0.0055 

0 . 0059 

0.0064 

100° 

0.0029 

0.0037 

0 . 0045 

0.0051 

0.0058 

0.0065 

0 . 007 1 

0.0077 

90° 

0.0034 

0.0043 

0.0052 

0.006 

0.0069 

0.0077 

0.0085 

0.0093 

80° 

0.004 

0.0052 

0 . 0063 

0.0073 

0.0083 

-  0.0093 

0.0102 

0.0112 

70- 

0 . 0048 

0.0065 

0.008 

0.0092 

0.0103 

Q.0115 

0.0124 

0.0133 

60° 

0.0059 

0.0084 

0.0103 

0.0119 

0.013 

0.014 

0.0148 

0.0156 

Conditions:  A  4-in.  journal  of  steel  6  in.  long  fitted  on  its  upper  side  with  a  gun 
metal  brass,  or  half  box,  embracing  somewhat  less  than  one-half  the  journal  circumfer- 
ence— the  chord  of  the  arc  contact  equal  to  3.92  in. — and  when  running  kept  at  a  constant 
temperature  of  90°  F. 

Table  15. — Coefficients  of  Friction  for  Bath  of  Lard  Oil — Load  Constant  at  100 
LB.  PER  Square  Inch  {Tower). 


Nominal 

load 

lb. 
per  sq.  in. 

Lb. 
520 
468 
415 
363 
310 

258 
205 

153 
100 


100  rev. 

105  ft. 

per  min. 


Speeds 


150  rev. 

157  ft. 

per  min. 


200  rev.  I  250  rev. 
209  ft.     I    262  ft. 


300  rev. 
314  ft. 


I  o . 0008 

O.OOII 

j  0.0012 

0.0013 

'  0.0015 

0.0014  0.0017 

0.0018  0.0021 

0.0023  0.003 

0.0036  0.0045 


per  mm.      per  mm.-  per  mm. 


o.ooi 

0.0013 

0.0014 

0.0016 

0.0017 

0.002' 

0.0025 

0.0035 

0.0055 


350  rev. 
366  ft. 
per  min, 


400  rev. 
419  ft- 


450  rev. 
471  ft. 


per  mm.!  per  mm. 


0.0012 

0.0013 

0.0014 

0.0015 

0.0014 

0.0015 

0.0017 

0.0018 

0.0015 

0.0017 

0.0019 

0.0021 

0.0017 

0.0019 

0.002 

0.0022 

0.0019 

0.0021 

0.0022 

0 . 0024 

0.0023 

0.0025 

0.0026 

0.0029 

0.0028 

0.003 

0.0033 

0 . 0036 

0.004 

0.0044 

0.0047 

0.005 

0.0063 

0 .  0069 

0.0077 

0.0082 

0.0017 

0.002 

0.0024 

0.0025 

0.0027 

0.0031 

0.004 

0.0057 

o . 0089 


Conditions:  A  4-in.  journal  of  steel  6  in.  long  fitted  on  its  upper  side  with  a  gun 
metal  brass,  or  half  box,  embracing  somewhat  less  than  one-half  the  journal  circumfer- 
ence— the  chord  of  the  arc  contact  equal  to  3.92  in. — and  when  running  kept  at  a  constant 
temperature  90°  F. 

Table  16. — Coefficients  of  Friction  for  Bath  of  Olive  Oil — Temperature  Con- 
stant {Tower). 


36 


BEARINGS  AND  THEIR  LUBRICATION 


Speeds 

load 

lb. 

per  sq.  in. 

100  rev. 

105  ft. 

per  min. 

150  rev. 

157  ft. 

per  min. 

200  rev. 

209  ft. 

per  min. 

250  rev. 
262  ft. 
per  min. 

300  rev. 
314  ft. 
per  min. 

350  rev. 

366  ft. 

per  min. 

400  rev. 
419  ft. 
per  min 

450  rev. 

471ft. 
per  min. 

Lb. 

573 
520 

415 
363 
258 
153 
100 

j 

i  0.00102 

1 
0.00055 

1  0 . 00093 

0 . 000S4 

0.00107    :    0.00139 

0.00162  ;  0.0020 

0.00277    0-00357 

0.00108 

0.00105 

0.00107 

0 . 0096 

0.00162 

0.00239 

0.00423 

0. 001 18 
O.OOII 5 
0.00II9 

O.OOII 

0  00178 
0.00267 
0.00503 

0.00126    0.00132 
0.00125    0.00133 
0.0013     :o. 00140 
0.00122   0.00134 
0.00195    0.00213 
0.003      ,0.00334 
0.00576   0.00619 

0.00139 
0.00142 
0.00149 
0.00147 
0.00227 
0.00367 
0 . 00663 

0.00148 
0.00158 
0.00155 
0.00243 
0.00396 
0.00714 

Conditions:  A  4-in  journal  of  steel  6  in.  long  fitted  on  its  upper  side  with  a  gun 
metal  brass,  or  half  box,  embracing  somewhat  less  than  one-half  the  journal  circumfer- 
ence— the  chord  of  the  arc  contact  equal  to  3 .  92  in. — and  when  running  kept  at  a  constant 
temperature  of  90°  F. 

Table  17. — Coefficients  of  Friction  for  a  Bath  of  Rape  Oil — Temperature  Con- 
stant {Tower). 


Nominal 

load 

lb, 

per  sq.  in. 


Speeds 


Actual 

load 
lb.  per 
sq.  in. 


Temper- 
ature 
Fahr. 


100  rev.      ISO  rev.  1  200  rev. 

105  ft.  157  ft.        209  ft. 

per  min.      per  min.     per  min. 

I 


250  rev.     300  rev. 

262  ft.    j    314  ft. 

per  min.  |  per  min. 


350  rev.  j  400  rev. 
366  ft.  I  419  ft. 
per  min.  ;  per  min. 


328 
310 
293 
275 
2S8 
20s 
153 
100 


582 
551 
520 
498 
4S8 
364 
272 
178 


0.0102 
o.oios 


0.C099 

0.0105 

0.0091 

0.0112 

O.OIOS 

0.009 

0.0099 


O.OT07 

0.0102 

0.0099 

0.0092 

O.OIOS 

0.0097 

0 .0091 

0.0095 

0.009S 

0.0088 

0.0087 

0.0085 

0.0096 

0.0102 

0.0109 

0.0122 

0.0098 
0.0099 
0.0097 
0.0103 
0.0084 
0.0078 
0.0105 
0.0133 


0 

0082 

0 

0085 

0 

0II9 

0 

0144 

0.0083 

O.OI 
O.OI2S 

0.0154 


Conditions:  A  4-in.  journal  of  steel  6  in.  long  fitted  on  its  upper  side  with  a  gun 
metal  brass,  or  half  box,  embracing  somewhat  less  than  one-half  the  journal  circumfer- 
ence— the  chord  of  the  arc  of  equal  to  contact  of  2.25  in. — and  when  running  kept  at  a 
constant  temperature  of  90°  F. 

Table  i8. — Coefficients  of  Friction  of  Rape  Oil  from  a  Pad  Under  the  Journal 

{Tower). 


COEFFICIENTS  OF  FRICTION 


37 


Actual 

Speeds 

Nominal 

load 

load 

100  rev. 

150  rev. 

200  rev. 

250  rev. 

300  rev. 

350  rev. 

400  rev. 

lb. 

lb. 

105  ft. 

157  ft. 

209  ft. 

262  ft. 

314  ft. 

366  ft. 

419  ft. 

per  sq.  in. 

per  sq. in. 

per  mm. 

per  mm. 

per  min. 

per  mm. 

per  mm. 

per  min. 

per  mm. 

Lb. 

Lb. 

258 

317 

0.0056 

0.0057 

0.0063 

0.0068 



205 

252 

0.0132 

0.0098 

0.007 

0.0077 

0.0082 

0.00S7 

100 

123 

0.0144 

0.0125 

0.0146 

0.0152 

0.0163 

0.0171 

0.0178 

Conditions:   Four-in.  journal,  6  in.  long;  chord  of  arc-3  1/4  in. 
Table  19. — Coefficients  of  Friction  for  Rape  Oil  Fed  by  Syphon  Lubricator  {Tower). 

COEFEICIENT  OF  FRICTION  OF  COLLAR  BEARINGS 

The  coefficient  of  friction  for  a  collar  bearing  is  much  higher  than  for  a 
journal  bearing  running  under  similar  conditions.  The  reason  lies  in  the 
difficulty  with  which  the  collar  bearing  is  lubricated.  There  is  little  or  no 
action  tending  to  introduce  oil  between  the  surfaces  as  in  journal  bearings. 


Load  in 

lb. 

total 

Speed  in  revolutions  per  minute 

50 

70 

90 

no 

130 

600 
1200 
1800 
2400 
2700 
3000 
3300 
^600 

0 . 0450 
0.0375 
0.0357 
0.0286 
0.0354 
0.0347 
0.0337 
0.0312 

0 . 0646 
0.0481 
0.0399 
0.0375 
0.0334 
0.0341 
0.0322 
0.0444 

0.0433 
0.0496 
0.0361 
0.0361 
0 . 0346 
0.0348 
0.0348 

0-0537 
0.0489 
0.0357 
0.0373 
0.0361 
0.0352 

0.0642 

0.0475 
0.0371 
0.0410 
0.0378 
0.0356 

Conditions:  Bearing  faces  12  in.  inside  diameter,  14  in.  outside  diameter,  one  of  soft 
steel,  the  other  of  gun  metal,  face  grooved  to  distribute  oil;  lubrication  sufficient  to  prevent 
seizing  with  a  mineral  oil. 

Table  20. — Coefficients  of  Friction  for  a  Collar  Bearing. 


The  coefficient  of  friction  for  collar  bearing  approaches  that  for  journal 
bearings  running  at  a  low  velocity  and  with  poor  lubrication,  and  tends  to  follow 
the  laws  for  friction  of  solids  rather  than  for  fluids.     It  may  be  tciken  as  from 


38 


BEARINGS  AND  THEIR  LUBRICATION 


0.04  to  0.05.  Experimental  data  tend  to  show  that  the  coefficient  is  independ- 
ent of  the  speed  of  rubbing  and  decreases  with  an  increase  of  load.  At  the 
same  time  this  type  of  bearing  is  only  usable  for  comparatively  low  unit  pres- 
sures and  for  comparatively  low  speeds,  unless  the  bearing  surfaces  are  separated 
by  a  film  of  oil  maintained  by  a  pressure,  or  forced  lubrication. 

Experiments  made  by  Mr.  Tower  and  reported  in  the  Proc.  Inst.  M.  E., 
1888,  page  179,  were  for  a  collar  bearing  14  in.  outside  diameter,  12  in.  inside 
diameter,  having  a  soft  steel  face  opposed  to  a  gun  metal  face.  Table  20  has 
been  compiled  from  these  experimental  results. 

FRICTION  OF  STEP  BEARINGS 

The  friction  of  step  bearings  is  about  the  same  as  of  collar  bearings.  Step 
bearings  are,  in  fact,  no  better  than  collar  bearings  except  that  they  are  generally 
of  smaller  diameters,  hence  have  a  lower  speed  and  are  therefore  less  liable  to 
overheating.     They  present  the  same  difficulties  as  regard  lubrication. 

It  is  very  difficult  to  assign  an  average  value  of  the  coefficient  for  a  step 
bearing.     However,  it  may  be  taken  as  o.oi  for  ordinary  machine  design. 

In  the  Proc.  Inst.  M.  E.,  1891,  page  114,  Tower  gives  results  of  experi- 
ments made  by  him  with  a  fiat  pivot  bearing  having  a  steel  step  opposed  to  a 
manganese  bronze-bearing  surface.  Table  21  gives  some  of  the  results  of 
those  experiments. 


Speed  in  revolutions  per  minute 

Load 

lb.   per 

Sc 

128 

194 

290 

353 

sq.  in. 

1 

Oil  drops 

Coeflf. 

Oil  drops 

Coeflf.  of 

Oil  drops 

Coeflf.  of 

Oil  drops 

Coeflf.  of 

Oil  drops 

Coeflf.  of 

per  min. 

of  fric. 

per  min. 

fric. 

per  min. 

fric. 

per  min. 

fric. 

per  min. 

fric. 

20 

20 

0.0196 

79 

0.0080 

196 

0.0102 

Con- 

0.0178 

Con- 

0.0167 

40 

26 

0.0147 

82 

0.0054 

stream 

0 . 006 I 

tinuous 

0.0107 

tinuous 

0.0096 

60 

40 

0.0167 

80 

O.OOS3 

200 

0.0051 

stream 

0.0078 

stream 

0.0073 

80 

so 

0.0181 

83 

0.0063 

200 

0.0045 

0.0064 

0.0063 

100 

54 

0.0219 

98 

0.0077 

200 

•  0.0044 

0.0056 

O.OOS7 

t20 

56 

0.0221 

84 

0.0083 

168 

0.0052 

0 . 0048 

O.OOS3 

140 

90 

84 

0.0093 
0.0113 

158 
168 

0.0062 
0.0068 

0.0046 
0 . 0044 

.00053 
O.OOS4 

160 

200 

260 

300 

seized 

' 

Conditions:    Steel  footstep  running  with  a  manganese  bronze  bearing   diameter  3  m.; 
lubrication  with  a  mineral  oi.  as  indicated. 


Table  21.— Coefficients  of  Friction  of  Pivot  Step  Bearing. 


COEFFICIFNTS  OE  FRICTION  39 

Another  series  of  experiments  was  made  with  white  metal  in  place  of  the 
l^ronze  bearing;  the  coefficient  of  friction  in  this  case  was  a  trifle  greater  than 
in  the  former,  but  the  difference  was  so  small  that  the  results  may  be  looked 
upon  as  practically  identical.  This  is  to  be  expected  when  the  friction  is  the 
friction  of  the  lubricant  and  not  of  the  metallic  surfaces. 

COEFFICIENT  OF  LATERAL  FRICTION 

The  endwise  floating  of  a  level  shaft,  provided  there  is  end  play  between 
the  bearings  and  journal  shoulders,  is  something  that  is  frequently  noticed. 
In  the  construction  of  electric  motors  and  generators  it  is  very  common  to 
leave  considerable  end  play  as  this  lateral  motion  of  the  shaft  aids  in  distribut- 
ing oil  throughout  the  bearings,  and  thus  improves  the  quality  of  the  lubrication. 
Anyone  who  has  had  occasion  to  touch  a  shaft  under  such  conditions  has  prob- 
ably been  surprised  at  the  very  small  amount  of  pressure  that  is  required  to 
make  the  shaft  move;  in  fact,  the  change  in  the  level  of  the  floor  by  shifting  of 
the  weight  of  a  person  walking  around  the  machine  is  often  sufficient  to  make 
the  shaft  move  endwise.  This  motion  can  be  likened  to  that  of  a  fine  pitch 
screw  rotating  and  moving  in  a  fixed  nut,  with  the  mating  thread  formed  by 
the  oil  film. 

Professor  Sweet  in  the  American  Machinist  for  February  16,  191 1,  page  315, 
gives  this  explanation  for  the  smallness  of  the  force  required  to  move  a  leveled 
and  well  lubricated  shaft  in  a  lateral  direction.  ''The  Straight  Line  engines 
were,  before  the  direct-connected  age,  built  with  a  quarter  of  an  inch  end 
play  in  the  main  bearings  and  crank.  When  running,  say  at  250  revolutions, 
with  2-ton  flywheels  and  4  3/4-in.  shafts,  it  was  easy  to  slide  the  wheels  and 
shafts  back  and  fourth  with  a  match.  My  explanation  was  this:  In  making, 
say  three  revolutions,  the  journals  were  sliding  about  4  ft.,  and  the  match 
moved  them  1/4  in.  Thus,  what  the  match  has  to  do  was  to  make  up  the 
difference  between  the  base  and  the  hypotenuse,  where  the  base  was  4  ft.  and 
the  perpendicular  1/4  in.,  which  was  very  little,  and  as  small  as  the  match  was, 
it  was  equal  to  do  that  litde  work. " 

John  Goodman,  in  a  paper  in  the  Proc.  Inst.  C.  E.,  Vol.  LXXXIX,  page 
424,  discusses  this  question  of  lateral  friction  in  an  experimental  apparatus, 
where  a  bronze  half  box  composed  of  an  alloy  of  i  part  of  tin  to  7  parts 
of  copper,  having  a  diameter  of  2  in.  and  a  length  of  4  in.,  was  run  in  contact 
with  a  manganese  steel  shaft.  At  a  circurnferential  velocity  of  200  ft.  per 
minute,  and  a  lateral  velocity  of  0.05  in.  per  minute,  the  lateral  coefficient  of 
friction  was  found  to  be  0.000021.  This  factor  is  given  here  more  as  a  matter 
of  interest  than  because  of  any  practical  value. 

In  general  the  lateral  coefficient  of  friction  varies  directly  as  the  lateral 
velocity,  and  inversely  as  the  circumferential  velocity. 


40  BEARINGS  AND  THEIR  LUBRICATION 

COEFFICIENT  OF  FRICTION  FOR  SURFACES  SLIDING  IN  STRAIGHT  PATHS 

The  coefficient  of  friction  for  surfaces  sliding  in  straight  paths  is  not  of  much 
importance,  as  ordinarily  guides,  ways  and  V's  of  machines  are  not  sub- 
jected to  great  pressures  and  do  not  move  with  high  velocities.  The  condition 
of  lubrication  is  that  of  unlubricated  or  very  poorly  lubricated  surfaces.  For 
metal  in  contact  with  metal  under  such  conditions  a  safe  factor  of  design  for 
coefficient  of  friction  is  0.3.  For  a  more  lengthy  discussion  of  the  subject 
reference  should  be  had  to  Section  I,  where  the  experimental  results  of  General 
Morin  are  presented. 

John  Aspinwall  in  Proc.  Inst.  C.  E.,  Vol.  CXXXIII,  page  13,  gives  results 
of  a  long  series  of  experiments  to  determine  the  coefficient  of  friction  of  locomo- 
tive slide  valves.  The  factors  are:  For  the  ordinary  i)-valve  with  its  face 
vertical,  0.068;  for  a  jD-phosphor  bronze  partially  balanced  valve  in  a  horizontal 
position,  0.0919;  for  an  unbalanced  cast-iron  valve,  0.0878. 


SECTION  III 
MATERIALS  FOR  BEARINGS 

Probably  nothing  in  connection  with  bearings  has  had  a  wider  discussion  or 
a  greater  amount  of  experimental  investigation  than  the  materials  from  which 
they  are  made.  All  of  this  implies  that  it  is  a  question  of  importance;  yet  its 
degree  of  importance  is  determined  to  a  great  extent  by  the  nature  and  quality 
of  the  lubrication  that  the  bearings  will  receive  in  use.  The  first  law  of  friction 
for  perfectly  lubricated  surfaces,  page  19,  is  that  the  coefficient  of  friction  for 
well  lubricated  surfaces  is  practically  independent  of  the  nature  of  those  sur- 
faces. Thus,  if  the  bearing  surfaces  can  be  kept  separated  by  a  film  of  oil 
the  wear  is  the  wear  of  the  oil  particles  and  not  of  the  surfaces  of  the  bear- 
ing and  its  journal.  The  principal  requisites  for  the  materials  are  that  they 
should  have  sufficient  mechanical  strength  to  sustain  the  pressures  to  which 
they  are  subjected,  and  that  they  should  tend  to  remain  smooth  and  to  conform 
to  each  other  under  the  rubbing  action,  thus  maintaining  the  conditions  neces- 
sary for  the  formation  of  the  oil  film. 

But  the  builder  of  machinery  has  no  control  over  the  lubrication  that  the 
bearings  of  his  machine  will  receive  after  they  have  been  put  into  service,  and 
it  is  well  known  they  are  often  abused  and  neglected.  Thus  the  conditions  of 
ordinary  lubrication  fall  into  the  class  of  poorly  lubricated  surfaces  and  the 
question  of  what  kind  of  material  to  use  assumes  considerable  importance, 
Metals  are,  of  course,  the  common  materials;  although  wood  and  fiber  have 
limited  applications. 

C.  B.  Dudley,  late  chemist  to  the  Pennsylvania  Railroad,  states  that  a  good 
bearing  metal  should  have  five  characteristics.  See  The  Journal  of  the  Franklin 
Institute,  February,  1892,  page  83.     These  are: 

1.  The  metal  must  have  strength  enough  to  sustain  its  load. 

2.  It  must  not  heat  rapidly. 

3.  It  must  work  well  in  the  foundry. 

4.  It  must  show  a  small  coefficient  of  friction. 

5.  It  must  give  a  large  amount  of  service  with  a  small  loss  of  metal  by  wear. 
To  these  five  we  might  add  a  sixth:  In  general  it  should  be  dissimilar  from 

the  metal  of  the  journal  with  which  it  is  to  run.  This  is  broadly  true,  except 
for  hardened  steel  in  contact  with  hardened  steel,  and  cast  iron  in  contact  with 
cast  iron.  There  seems  to  be  no  entirely  satisfactory  reason  for  this  dissimi- 
larity except  it  is  claimed  that  wear  is  more  rapid  and  friction  greater  when  the 

41 


42  BEARINGS  AND  THEIR  LUBRICATION 

journal  and  its  bearing  are  made  from  the  same  metal.  It  is  possible  that 
adhesion  has  something  to  do  with  it. 

On  page  5  it  is  pointed  out  that  two  metallic  surfaces  in  intimate  contact 
may  be  held  together  by  a  considerable  force  that  we  call  adhesion  or  molecular 
attraction. 

The  ordinary  combinations  of  metals  used  for  spindles  or  shafts  and  their 
bearings  are  as  follows: 

Hard  steel  journals  in  contact  with  hard  steel  bearings,  used  for  high  speeds 
and  moderate  pressures. 

It  calls  for  excellent  workmanship,  as  both  journal  and  bearing  must  be 
accurately  ground  and  carefully  alined.  Under  such  conditions  the  service 
rendered  is  good. 

Hard  steel  journalswith  soft  steel  or  wrought  iron  bearings.  This  combina- 
tion is  found  in  some  presses  where  a  hardened  steel  toggle  pin  fits  into  soft 
steel  seats.  It  is  generally  considered  to  give  good  service  for  high  pressures 
and  low  speeds. 

Hard  steel  journals  with  bronze  bearings.  This  is  a  very  common  com- 
bination in  machine  tools. 

Hard  steel  journals  with  cast  iron  bearings.  This  combination  is  found  to 
a  limited  extent  in  machine  tools.  The  use  of  cast  iron  as  a  bearing  metal  has 
had  a  peculiar  history,  with  strong  advocates  and  just  as  strong  opponents. 
It  is  of  such  importance  that  is  is  treated  of  in  detail  on  page  43. 

Soft  steel  journals  with  bronze  bearings.  This  combination  is  the  most 
commonly  used  of  any  among  machine  tools  and  high-grade  machinery  and 
gives  good  results. 

Soft  steel  journals  with  babbitt  bearings.  This  combination  is  the  one 
most  commonly  used  in  machinery  in  general. 

Soft  steel  journals  with  cast  iron  bearings.  This  combination  is  used  to 
some  extent  in  machine  tools  and  to  a  much  greater  extent  in  medium  and  light 
weight  machinery  where  the  bearing  pressures  are  small  and  the  speeds  low. 

Soft  steel  journals  and  lignum-vitae  bearings.  This  combination  has  been 
used  for  line  shaft  bearings  in  the  past.  Now  it  is  used  for  the  steps  of  vertical 
water  turbines.  Tail  shaft  bearings  of  steamships  are  a  bronze  sleeve  against 
lignum  vitae. 

Cast  iron  with  cast  iron.  This  combination  is  a  great  exception  to  the 
disuse  of  like  metals.  It  is  employed  for  the  step  bearings  of  vertical  steam 
turbines,  to  a  limited  extent  for  spindles  and  bearings  of  machine  tools,  is  com- 
monly used  for  the  piston  rings  and  cylinders  of  all  kinds  of  reciprocating 
engines,  has  been  used  in  the  crossheads  and  slides  of  steam  engines  and  air 
compressers,  and  is  the  combination  commonly  found  in  rectilinear  slides  and 
guides  of  all  kinds  of  machinery. 


MATERIALS  FOR  BEARINGS  43 

WEAR  OF  BEARING  METALS 

The  wear  of  bearing  metals  is  of  importance  in  railroad  practice  for  it 
represents  an  enormous  loss  of  valuable  metal  each  year  in  the  amount  rubbed 
off  and  lost.  In  electrical  machinery,  such  as  motors  and  generators,  the 
wearing  of  the  bearing  tends  to  decenter  the  armatures  and  rotors  and  thereby 
unequalize  the  air  gap  to  the  disadvantage  of  the  operating  of  the  machine.  In 
any  class  of  machinery  where  the  alinement  of  the  shafts  is  of  importance,  the 
wear  of  the  bearing  is  likewise  of  importance,  for  this  determines  the  position 
of  the  shafts. 

Dr.  Dudley  in  the  Journal  of  the  Franklin  Institute^  March,  1892,  page  169, 
lays  down  three  elements  in  the  wear  of  bearing  metals.     These  are: 

1.  That  metal  which  will  suffer  the  most  distortion  without  rupture  will 
wear  best.  Or,  stated  a  little  differently:  That  metal  which  has  the  greatest 
percentage  of  elongation  will  wear  the  best. 

2.  With  a  satisfactory  elongation  an  increase  in  tensile  strength  will  add  to 
the  wearing  property  of  the  metal. 

3.  With  equal  percentages  of  elongation  and  equal  tensile  strengths  the 
finer  the  granular  structure  of  the  metal  the  longer  will  it  wear. 

The  idea  of  wear  is  the  rubbing  off  or  rupturing  of  minute  particles.  Thus 
a  metal  which  has  the  greatest  tensile  strength  to  resist  rupture  and  the  property 
of  withstanding  the  greatest  amount  of  distortion  without  rupture  and  a  fine 
granular  structure,  so  that  the  particles  that  are  rubbed  off  are  minute,  can  be 
considered  as  having  the  greatest  resistance  to  wear.  There  are  few  experi- 
mental data  to  support  the  relationship  between  these  variables,  although  the 
conclusion  seems  to  be  well  founded  that  an  increase  of  tensile  strength  at  the 
expense  of  elongation  is  decidedly  to  the  detriment  of  the  wearing  qualities. 

In  following  paragraphs  dealing  with  bronzes  this  question  of  wear  will  be 
again  considered  in  quantitative  factors. 

CAST  IRON 

It  is  emphasized  by  its  advocates  and  conceded  by  its  opponents  that  cast 
iron  under  certain  conditions  will  take  on  a  hard  glaze  that  makes  a  splendid 
wearing  surface.  This  glaze  is  so  hard  that  it  will  resist  the  cutting  edge  of  a 
scraper.  Furthermore,  cast  iron  is  porous  and  readily  absorbs  oil,  thus  to  a 
degree  becoming  self-lubricating.  It  is  granular  in  structure,  in  wear  it  rubs 
off  in  minute  particles.  It  can  be  obtained  in  varying  degrees  of  hardness, 
even  chilled.  Other  advantages  are  the  ease  with  which  it  is  handled  in  the 
foundry  and  machine  shop,  and  its  cheapness. 

The  objection  to  it  is  summed  up  by  saying  that  it  is  a  'treacherous"  metal. 
After  glazing  the  bearing  is  very  permanent  and  has  great  resistance  to  cutting; 


44  BEARINGS  AND  THEIR  LUBRICATION 

but  before  glazing,  if  cutting  is  once  started  it  takes  but  a  very  short  time  to 
ruin  the  bearing.  With  an  assurance  of  adequate  lubrication,  the  case  against 
cast  iron  as  a  bearing  metal  would  at  once  break  down. 

Professor  Sweet  has  long  been  an  earnest  advocate  of  the  cast-iron  box,  as 
frequent  articles  and  discussions  in  the  American  Machinist  show.  William 
Sellers  &  Company,  of  Philadelphia,  has  been  furnishing  cast-iron  lineshaft 
bearings  for  years. 

R.  K.  LeBlond  in  the  American  Machinist,  March  23,  191 1,  page  537,  gives 
experiences  with  four  experimental  lathes  fitted  with  spindles  and  bearings  as 
follows: 

1.  Hardened  steel  spindle  with  cast-iron  boxes. 

2.  Soft  steel  spindle  with  babbitt  boxes. 

3.  Hardened  steel  spindle  with  bronze  boxes. 
.     4.  Soft  steel  spindle  with  bronze  boxes. 

The  soft  steel  spindles  were  of  60-carbon  crucible  steel;  the  bronze  was 
made  to  the  specifications  of  the  Pennsylvania  Railroad  Company.  After  the 
end  of  some  8  years'  service  and  treatment  as  far  as  possible  identical  for  all 
four  lathes,  it  was  found  that  their  rating  as  regards  absence  of  wear  and  general 
satisfaction  was  in  the  order  as  given  above;  that  is,  the  hardened  steel  spindle 
with  cast-iron  boxes  was  the  best  combination.  Both  spindle  and  boxes  were 
in  as  good  condition  as  when  placed  in  the  lathe,  and  from  all  appearances  and 
tests  showed  absolutely  no  wear. 

Mr.  LeBlond  says  further:  "The  dry  bronze  box  with  soft  spindle  will  stand 
more  abuse  than  a  dry  cast-iron  box  and  soft  spindle;  but  a  dry  cast-iron  box 
and  hard  spindle  will  withstand  more  abuse  than  the  bronze."  And  again: 
''The  hardened  steel  spindle  and  cast-iron  box  will  stand  as  much  neglect  as 
any  combination  of  metals,  has  a  much  longer  life,  will  retain  its  accuracy  for  an 
indefinite  period,  will  stand  intermittent  cuts  or  series  of  blows  which  would 
peen  out  and  loosen  babbitt,  and,  from  our  experience,  is  positively  the  best  and 
most  lasting  bearing  ever  put  into  a  machine  tool." 

The  experience  of  the  John  Steptoe  Shaper  Company  is  shown  by  the  follow- 
ing quotations  taken  from  an  article  of  Professor  Sweet's  printed  in  the  American 
Machinist. 

"  We  have  been  using  cast-iron  bearings  on  all  our  machines  since  July,  1909. 
Since  that  time  we  have  turned  out  over  500  machines,  and  we  have  yet  to  re- 
ceive the  first  complaint  with  regard  to  the  bearings  running  hot;  but  we  have 
made  proper  provision  for  the  distribution  of  the  oil  over  the  bearings.  We 
have  chased  spiral  oil  grooves  in  the  shafts  and  have  provided  the  bearings 
with  ring  oilers  so  that  they  are  constantly  flooded  with  oil.  We  believe  that 
the  cast-iron  box  is  the  best  bearing  that  we  have  ever  put  in  our  machine." 

Referring  to  Professor  Sweet's  own  work,  he  says:  ''In  the  artisan  lathe  J 


MATERIALS  FOR  BEARINGS  45 

put  a  ground  steel  shaft  in  a  solid  cast-iron  box.  The  outside  of  the  box  fitted 
in  a  conical  piece  with  a  nut  at  each  end,  but  was  not  split,  so  when  they  first 
became  loose  the  boxes  wdll  compress  enough  without  splitting,  and  when  not 
enough  they  can  be  split." 

Cast  iron  bearings  have  been  extensively  used  in  shoe  machinery  for  many 
years,  and  in  textile  machinery.  One  of  the  most  troublesome  places  in  the 
machine  shop  is  in  countershaft  bearings.  It  has  long  been  accepted  practice 
in  many  places  to  bush  countershaft  hangers  and  loose  pulleys  with  cast  iron 
when  the  original  bearings  wore  out. 

Coleman  Sellers,  Jr.,  writes  of  the  practice  of  his  firm  in  regard  to  the 
use  of  cast-iron  bearings  and  spindles  as  follows: 

"  William  Sellers  &  Company  have  always  believed  in  cast-iron  bearings  where  the  condi- 
tions of  feed  and  pressure  are  not  excessive  and  where  efficient  lubrication  can  be  insured. 
Under  such  conditions  we  believe  that  a  film  of  oil  separates  the  metallic  surfaces  so  that  they 
do  not  come  in  contact.  Where  the  load  or  speed  are  such  that  the  film  of  oil  is  likely  to  be 
disturbed  then  it  becomes  necessary  to  use  some  material  in  the  bearing  not  so  likely  to  adhere 
or  'cut.'  It  has  always  been  our  practice  to  use  cast-iron  boxes  in  our  own  shafting,  and  we 
have  had  cases  where  boxes  have  run  for  30  years,  not  only  without  cutting  but  without 
appreciable  wear. 

We  have  built  many  machines  with  cast-iron  spindles  and  cast-iron  bearings  and  with  steel 
spindles  in  cast-iron  bearings  both  with  excellent  results.  We  realize  that  such  bearings  must 
be  thoroughly  lubricated.  They  will  not  run  well  if  dry.  We  have  carried  heavy  vertical 
shafts  on  cast-iron  steps  with  perfect  results;  but  care  has  been  taken  to  make  these  steps  so 
large  that  the  unit  pressure  would  be  low  and  the  oil  circulation  was  made  very  effective. 
Boring  mill  tables,  planer  tables,  boring  mill  center  spindles  and  boring  mill  steps  are  all 
emaples  of  cast-iron  running  on  cast-iron,  and  it  is  well  understood  that  under  proper  condi- 
tions of  load,  speed  and  lubrication  such  bearings  are  perfectly  satisfactory." 

In  the  Proc.  Inst.  C.  E.,  Vol.  LXXXV,  in  a  discussion  by  Dr.  Goodman,  it 
is  stated  that  on  the  Brighton  Railroad  in  England  cast  iron  eccentric  straps 
working  on  cast-iron  hubs  have  run  at  least  100,000  miles  without  having  been 
taken  up. 

These  few  quotations  are  sufficient  to  show  the  esteem  in  which  cast  iron  is 
held  as  a  bearing  metal,  provided  it  is  used  under  advantageous  conditions. 
The  common  prejudice  against  it  is  unfounded,  but  one  must  not  lose  sight  of 
the  fact  that  a  cast  iron  bearing  will  not  run  dry. 


SOFT  ALLOYS  OR  WHITE  METALS 

Babbitts  are  white  bearing  metals  whose  most  common  constituents  are 
tin,  lead,  and  antimony.  As  distinguished  from  a  brass  or  bronze,  babbitt 
metal  can  be  melted  in  an  ordinary  ladle.  This  is  one  of  the  governing  rea- 
sons for  its  use  as  a  lining  for  journal  bearings.     It  is  easily  melted,  easily 


46  BEARINGS  AND  THEIR  LUBRICATION 

poured  and  run,  easily  anchored  in  place  in  its  shell,  has  good  anti-friction 
properties,  and  may  be  poured  around  the  journal  with  which  it  is  to  run 
(although  this  is  not  approved  practice).  Thus  the  babbitting  process  is  the 
easiest  method  known  for  obtaining  well  alined  bearings. 

B.  R.  Tompkins,  writing  in  the  Mechanical  News  for  January,  1891,  says: 
"For  slow  running  journals,  where  the  load  is  moderate,  almost  any  metal 
that  may  be  conveniently  melted  and  will  run  free  will  answer  the  purpose. 
For  wearing  properties,  with  a  moderate  speed,  there  is  probably  nothing 
superior  to  pure  zinc,  but  when  not  combined  with  some  other  metal  it  shrinks 
so  much  in  cooling  that  it  cannot  be  held  firmly  in  the  recess,  and  soon  works 
loose;  and  it  lacks  those  anti-friction  properties  which  are  necessary  in  ordar 
to  stand  high  speed. 

''For  line  shafting  and  all  work  where  the  speed  is  not  over  300  or  400 
r.  p.  m.  an  alloy  of  8  parts  zinc  and  2  parts  block  tin  will  not  only  wear  longer 
than  any  composition  of  its  class,  but  will  successfully  resist  a  heavy  load. 
The  tin  counteracts  the  shrinkage,  so  that  the  metal,  if  not  overheated,  will 
firmly  adhere  to  the  box  until  it  is  worn  out.  But  this  mixture  does  not  possess 
sufficient  anti-friction  properties  to  warrant  its  use  for  fast  running  journals. 

"Among  all  the  soft  metals  in  use  there  is  none  that  possesses  greater  anti- 
friction properties  than  pure  lead;  but  lead  alone  is  impracticable,  for  it  is  so 
soft  that  it  cannot  be  retained  in  the  recess.  But  when  by  any  process  lead  can 
be  sufficiently  hardened  to  be  retained  in  the  boxes  without  materially  injuring 
its  anti-friction  properties,  there  is  no  metal  that  will  wear  longer  in  light  fast 
running  journals.  With  most  of  the  best  and  most  popular  anti-friction  metals 
in  use  and  sold  under  the  name  of  'babbitt  metal,'  the  basis  is  lead. 

"Lead  and  antimony  have  the  property  of  combining  with  each  other  in 
all  proportions  without  impairing  the  anti-friction  properties  of  either.  The 
antimony  hardens  the  lead,  and  when  mixed  in  the  proportion  of  80  parts  of 
lead  by  weight  with  20  parts  antimony,  no  other  known  composition  of  metals 
produces  greater  anti-friction  or  wearing  properties,  or  will  stand  a  higher 
speed  without  heat  or  abrasion.  It  runs  free  in  its  melted  state,  has  no  shrink- 
age and  is  better  adapted  to  light  high  speed  machinery  than  any  other  known 
metal.  Care,  however,  should  be  manifested  in  using  it,  and  it  should  never  be 
heated  beyond  a  temperature  that  would  scorch  a  dry  pine  stick. 

"Many  different  compositions  are  sold  under  the  name  of  babbitt  metal. 
Some  are  good,  but  more  are  worthless.  But  very  little  genuine  babbitt  metal 
is  sold  that  is  made  strictly  according  to  the  original  formula,  most  of  the 
metals  sold  under  that  name  are  the  refuse  of  type  foundries  and  other  smelting 
works,  melted  and  cast  into  fancy  ingots  with  special  brands,  and  sold  under 
the  name  of  babbitt  metal, " 

It  is  difficult  to  determine  the  exact  formula  used  by  the  original  discoverer, 


MATERIALS  FOR  BEARINGS 


47 


Babbitt.     Tin,  copper  and  antimony  were  the  ingredients,  and  from  the  best 
sources  of  information  the  original  proportions  in  per  cent,  were  as  follows: 

Tin     =89.3  or  83.3  or  89. 1. 

Copper=   3.6  or    8.3  or    3.7. 

Antimony=   7 .  i  or    8 . 3  or    7.4. 

This  metal,  when  carefully  prepared,  is  probably  one  of  the  best  metals  in 
use  for  lining  boxes  that  are  subjected  to  a  heavy  weight  and  wear;  but  for  light 
fast  running  journals  the  copper  renders  it  more  susceptible  to  friction. 

BRONZE  BEARING  METALS 

Next  in  importance  to  the  babbitt  is  the  bronze  series  of  bearing  metals.  A 
bronze  is  commonly  taken  to  mean  an  alloy  of  copper  and  tin  with  small 
amounts  of  other  constituents.  A  brass  is  an  alloy  in  which  copper  and  zinc 
are  the  principal  constituents. 


Detail  showing  Joint      Cut  Oue  Lug 
of  Packing  Ring.  '"  ««ch  Seg- 

meut  thus  for 
Liftiug. 

FIG.    7.       BEARING    METAL   APPLIED    TO   A    STEAM   ENGINE    PISTON. 

In  contrast  to  the  softer  metals,  they  must  be  melted  in  a  crucible  and 
cannot  be  manipulated  in  the  easy  manner  in  which  the  soft  metal  linings  are 
made.  As  in  the  case  of  babbitt  metals  there  are  a  multitude  of  bronzes  on  the 
market. 

In  an  intermediate  position  between  the  babbitt  series  and  the  bronze 
series  is  a  metal  made  by  A.  Allan  &  Son,  known  as  Allan  "red  metal."  It  is 
composed  of  50  per  cent,  copper  and  50  per  cent.  lead.  It  is  a  soft  metal,  red  in 
color,  will  withstand  a  temperature  of  575°  F.  without  injury,  may  be  poured 


48 


BEARINGS  AND  THEIR  LUBRICATION 


into  a  lining  shell,  but  not  around  a  metal  mandrel,  and  is  extensively  used  for 
the  bearing  rings  of  reciprocating  steam  engine  pistons.  Fig.  7  shows  the  way 
in  which  it  is  used  for  this  purpose.     A  similar  metal  is  ''Plumbic  Bronze." 

In  the  preceding  paragraphs  it  was  pointed  out  that  lead  is  one  of  our  most 
useful  bearing  metals  when  used  with  others  to  make  an  alloy.  Two  theories 
have  been  advanced  as  to  the  relative  action  of  the  constituents.  The  one  is 
that  the  harder  metal  which  forms  a  net  work  around  particles  of  the  softer 
metal  is  the  important  one.  That  is,  it  is  the  harder  metal  that  takes  the  wear 
and  determines  the  frictional  value  of  the  alloy;  the  softer  metal  merely  fills  up 
the  spaces  throughout  the  network  of  the  harder  matrix,  and  wears  away  more 
rapidly  in  order  that  the  matrix  may  always  be  in  contact  with  the  journal. 

The  other  theory  is  that  the  softer  metal  is  the  one  that  determines  the  anti- 
friction qualities  of  the  alloy.  This  is  the  explanation  that  is  more  generally 
accepted,  and  in  view  of  the  fact  that  we  know  that  lead  is  a  most  valuable  con- 
stituent in  these  alloys,  we  are  still  further  led  to  believe  that  it  is  the  soft  metal 
that  is  determinative. 

Andrew  Allan,  of  the  firm  of  A.  Allan  &  Son,  was  the  first  manufacturer  of 
bronzes  in  America  to  point  out  the  value  of  lead  in  large  proportions  as  a  valu- 
able constituent  of  copper-lead-tin  bearing  metal.  He  has  made  such 
alloys  since  1876.  He  has  standard  bronzes  with  different  proportions  of  cop- 
per, lead  and  tin,  for  different  services,  as  indicated  below: 


Number 

Copper  in 
per  cent. 

Lead  in 
per  cent. 

Tin  in 
per  cent. 

Uses 

2 

66 

25 

9 

For    severe    service    as    mill    pinion 
bearings. 

4 

59-5 

35 

5-5 

For  locomotive  brasses. 

5 

50-5 

45 

4-5 

For  passenger  and  freight  car  brasses. 

6 

48.75 

48.75 

2-5 

For  service  similar  to  that  of  hard 
babbitt. 

Name  of  metal 

Copper 

Tin 

Lead 

1 

Anti- 
mony 

Zinc 

Iron 

Camelia  metal 

70.2 
1.6 

4.2s 
98.13 

X4.75 

10.2 

0.55 

White  metal 

87.92            ' 
84.87 

I.IS 

12.08 
15-1 

Car  brass  lining 

trace 
9.91 

Salgee  metal 

4.01 

85.57 

■ 

Table  22. — ^Analyses  of  Common'  Bearing  Metals  {Dudley). 


MATERIALS  FOR  BEARINGS 


49 


Name  of  metal 

Copper 

i        Tin        1       Lead       |      And- 

Zinc     !      Iron 

14-38 
Graphite — 

67-73 
none 

80.69 

14-57 
possible 

trace 

12.4 
— trace 

S-i 
83-55. 
on,  copper, 

78.44 

0.31 

15. c6 

12.52 

16.73 
18.83 

tcrmined 

Carbon  bronze 

7 "C   A^ 

9.72 
Carbon— 

9.6 

Phosphorus 

2.37 

77.83 
92.39 

trace 

Delta  metal 

Llagnolia  metal 

16.45 
zinc  and  pc 

19.6 

American  anti-friction   metal.... 

59 

7C.8 

Traces  of  ir 
bismuth. 

2.16 
9.2 
10.6 
9.58 

Manganese 

10.98 
Phosphorus 

ssibly 

0.98 
38.4 

0.65 

Graney  bronze 

I.Ianganese  bronze 

90.52 
81.24 

Ajax  metal 

—none 

7.27 
or  arsenic 

88.32 

0.37 
11.93 

Harrington  bronze 55-73 

0.97 

42.67 
trace 

0.68 

84.33 
94-4 
9.61 
—  0.94 

1 5 



14-38 
6.03 

Hard  lead 

Phosphor  bronze 

79-17 
76.8 

10.22 
Phosphorus 

8 
Phosphorus 

Ex.  B    metal     . 

— 0.  2 

Table  22  {Continued). — Analyses  of  Common  Bearing  Metals  (Dudley). 


ANALYSES  AND  PHYSICAL  PROPERTIES  OF  BEARING  METALS 


Dr.  Dudley,  in  The  Journal  of  the  Franklin  Institute,  February,  1892, 
page  87,  gives  a  number  of  analyses  of  common  bearing  metals,  mostly  bronzes 
as  determined  in  the  laboratory  of  the  Pennsylvania  Railroad.  Table  22 
has  been  compiled  from  the  results  there  given. 

G.  H.  Clamer  also  in  The  Journal  of  the  Franklin  Institute  gives  a  large 
number  of  analyses  of  bearing  metals  from  which  Tables  23-26  have  been 
compiled.  These  are  of  value  as  showing  a  wide  variety  of  both  American 
and  European  practice. 


50 


BEARINGS  AND  THEIR  LUBRICATION 


Tin 

Copper 

Antimony 

References 

96 

4 

8 

Quoted  by  Thurston.     Used  for  ordinary  bearings. 

90 

2 

8 

Quoted  by  Thurston.  Quoted  by  Hiorns  for  bearings  heavily  loaded , 
used  by  Russian  railroads  for  car  bearings. 

83.8 

3-7 

7-4 

Quoted  by  Thurston  and  BoUey,  as  Karmarsch  metal.  Used  in 
France  in  naval  constructions. 

87.5 
87 

12    5 

Quoted  by  Thurston,  Karmarsch  metal. 

6 

7 

Quoted  by  Hiorns,  for  bearings  heavily  loaded. 

8S 

5 

10 

Quoted  by  Ledebur  and  Hiorns  as  Jacoby  metal  for  light  pressure. 

83.33 

S-SS 

II .  II 

Used  for  car  bearings,  "Compagnies  de  I'Est,  P.  L.  M.,  Quest,"  etc. 

8] 

6 

II 

Quoted  by  Ledebur.     Used  by  Berlin  railroads. 

82 

6 

12 

Quoted  by  Ledebur.  Used  by  Orleans  and  the  Western  Austrian 
railroads. 

82 

8 

10 

Bearings  for  valve  rods  and  eccentric  collars,  "  Compagnie  du  Nord." 

81 

5 

14 

Quoted  by  Hiorns,  for  very  hard  bearings. 

80 

10 

10 

Quoted  by  Thurston.     Used  by  Swiss  railroads. 

78. s 

10 

II-5 

Quoted  by  Thurston.     Used  by  Russian  railroads. 

76.7 

7.8 

15. 5 

Quoted  by  Ledebur  and  Thurston  as  English  alloy. 

76 

7 

17 

Quoted  by  Hiorns,  for  bearings  lightly  loaded. 

Quoted  by  Thurston  and  BoUey,  as  Karmarsch  metal. 

73 

9 

18 

Quoted  by  Thurston  and  Hiorns,  for  light  pressures. 

71.4 

21.4 

7.2 

Quoted  by  Thurston.     Karmarsch  metal. 

71 

S 

24 

Thurston  standard  white  metal.  Used  by  the  P.  L.  M.  Company 
for  packing  of  valves  and  eccentric  collars. 

67 

22 

II 

Quoted  by  Thurston.    Used  by  the  Great  Western  Railway  (England) 

67 

n 

22 

French  state  railroads. 

33-3 

22.2 

44.5 

Quoted  by  Hiorns.     Dewrance  metal  for  locomotives. 

12 

4 

82 

Quoted  by  Hiorns,  for  very  hard  bearings. 

Table  23. — Alloys  of  Tin,  Copper  and  Antimony  (Clamcr). 


Lead 

Tin 

Antimony 

85 

.s 

84 

16 

80 

12 

8 

77.7 

5-9 

16.8 

76 

14 

10 

73 

12 

15 

70 

20 

10 

68 

IS 

17 

60 

20 

20 

42 

46 

12 

42 

42 

16 

37 

38 

25 

References. 


Soft  alloy  quoted  by  Hiorns.     Quoted  also  by  Dudley. 

Quoted  by  Ledebur.     For  slow  revolving  pulleys. 

Used  by  Eastern  Railroad  (France)  for  metallic  packings. 

Quoted  by  Thurston,  as  being  the  composition   of  Magnolia  and 

Tandem  metals. 

Used  for  metallic  packings  by  the  Orleans  and  P.  L.  M.  Railroads 

Metallic  packings  of  piston  rods.     Northern  Company  (France). 

Metallic  packings  of  eccentric  collars.     French  state  railroads. 

Graphite  ( ?)  metal  analyzed  by  Dudley. 

Quoted  by  Ledebur  and  used  for  railroad  bearings. 

Quoted  by  Hiorns.     Hoyles  metal. 

Quoted  by  Ledebur.     Used  for  journal  boxes,  French  state  railroads. 

Quoted  by  Thurston.      Italian  railroad  companies. 


Table  24. — Alloys  of  Lead,  Tin  and  Antimony  (Clamer). 


MATERIALS  FOR  BEARINGS 


51 


Copper 


Tin 


Zinc 


References 


86 

14 



84 

14 

! 

2 

83 

IS 

3 

82 

x8 

82 

15 

! 

3 

82 

16 

2 

80 

18 

! 

2 

78.7 

6 

3 

IS 

78 

20 

2 

77-4 

15 

6 

7 

S8 

28 

14 

57 

14 

29 

56 

28 

16 

6 

14 

i 

80 

S-S 

17 

S 

77 

Quoted  by  Thurston  for  locomotive  bearings. 

Quoted  by  Thurston.     Italian  railroads. 

Quoted  by  Bolley  as  a  hard  alloy. 

Quoted  by  Thurston  for  locomotive  bearings.  , 

Used  by  French  State  railroads  for  pieces  subiectcd  to  alternative 

friction. 
"Lafond"  alloy  for  heavily  loaded  bearings. 
Quoted  by  Ledcbur  and  Thurston.     Car  bearings  of  the  "  Compagnie 

du  Nord." 
Used  by  Orleans  Railroad  for  valve-rod  bearings. 
Quoted  by  Ledebur  for  locomotive  bearings.     Used  by  French  State 

railroads  for  pieces  subjected  to  circular  friction. 
Quoted  by  Thurston  and  Bolley  as  "Lafond"  alloy. 
Quoted  by  Bolley  for  friction  upon  cast  iron. 
Quoted  by  Bolley  and  Thurston  for  car  bearings. 
Quoted  by  Haswell  as  a  hard  bronze  for  bearings. 
Quoted  by  Thurston  as  "Margraflf"  alloy. 
Quoted  by  Hiorns  for  bearings  for  propclloi  shafts. 
Quoted  by  Thurston,  "Fenton"  alloy. 
Quoted  by  Hiorns  and  Thurston.     "Fenton"    alloy   for  locomotive 

and  car  bearings. 
Quoted  by  Ledebur,  for  high-speed  horizontal  shafts. 


Table  25. — Alloys  of  Copper,  Tin  And  Zinc  {Clamer). 


Copper 

Tin 

Lead 

Zinc 

Iron 

Anti- 
mony 

Phos- 
phorus 

References 

10            65 

25 

Bearings  for  locomotives  and  tenders 

74                0 . 1   1        7 

95 

5 
83.3 

76.2 
52 

-■JO 

Quoted  by  Thurston  and  Ledebur      Loco- 

8 
7.6 

I7-S 
46 
15 
32 

8 
3 

0.  T 

25 
60 

motive  bearings. 
Quoted  by  Bolley  and  Ledebur.      Bearings 
for  engines. 
Quoted  by  Ledebur   as   "Pierrot"  metal. 

"  Beugnot"  white  bronze  used  in  France 

in  naval  constructions. 
"White  bronze  used  for  ship  engines. 
Dunnlevic  and  Jones  metal. 
Quoted  by  Ledebur,  "Kniess"  metal. 
Used  by  the  western  railroads  (France)  for 

piston  rods  and  eccentric  packings. 

8.3 
5.6 

3.8 

1.6 

0.4 

5 

3 

Table  26. — Miscellaneous  Alloys  {Clamer). 


An  increasing  amount  of  attention  is  being  given  to  the  physical  properties 
of  bearing  metals;  the  tensile  and  compressive  strengths,  elongation,  reduction 
of  area  and  Brinell  hardness  number.  It  is  recognized  that  a  suitable  chemical 
analysis  alone  is  not  sufficient  to  ensure  physical  characteristics  that  will  give  a 
good  alloy  in  service.     From  the  records  of  tests  in  the  laboratory  of  the  Pennsyl- 


52  BEARINGS  AND  THEIR  LUBRICATION 

vania  Railroad  Table  27  has  been  arranged.  This  gives  the  physical  proper- 
ties of  many  commercial  alloys. 

Table  28  give  representative  chemical  analyses  of  many  of  these  same  alloys, 
but  the  two  tables  must  not  be  connected  directly.  For  small  variations  in  the  con- 
stituents of  such  metals  are  common,  while  the  effect  of  such  variations  cannot  be 
stated  definitely  it  is  possible  that  small  amounts  of  some  elements  may  have  a 
very  decided  influence  upon  the  physical  properties  of  the  alloy  that  they  enter. 

This  has  led  some  firms  to  insist  that  nothing  but  ingot  metal  of  known  com- 
position shall  be  used  either  in  melting  babbitts  or  in  charging  crucibles  for 
bronze  mixtures.  In  no  case  can  scrap  metal  or  the  fins,  or  sprues,  or  defective 
castings  from  previous  pourings  be  remelted.  All  such  metal  is  sent  to  a  refining 
room  where  it  is  analyzed  and  used  to  form  the  standard  mixtures  that  are  issued 
in  the  form  of  ingots  to  the  babbitting  room  and  bronze  foundry.  This  case 
is  exercised  to  keep  the  mixtures  uniform,  standard  and  under  control. 

In  preparing  the  ingot  metal,  chemical  tests  alone  are  not  relied  upon, 
however,  but  the  physical  properties  are  likewise  investigated.  Considerable 
weight  is  given  to  the  Brinell  hardness  number,  although  in  some  place  the 
sclerescope  hardness  number  is  the  one  found  and  considered  instead.  In 
this  connection  it  is  interesting  to  note  the  same  bronzes  will  give  a  sclere- 
scope hardness  number  as  high  as  high-grade  alloy  steel.  This  is  probably 
due  to  the  elasticity  of  the  metals,  for  the  sclerescope  measures  the  recupera- 
tive power  of  the  specimen  tested  when  a  permanent  deformation  is  caused 
by  the  impact  of  the  hammer.  Sclerescope  readings  for  different  metals 
should  not  be  compared. 

Comparisons  of  Brinell  and  sclerescope  hardness  numbers  for  the  same 
specimens  indicate  that  there  is  some  relationship  between  them,  though  just 
what  has  not  been  accurately  determined. 

In  testing  babbitts  for  hardness,  and  in  fact  other  physical  properties,  care 
must  be  exercised  to  use  specimens  of  the  same  general  size  and  cast  under 
similar  conditions.  Babbitt  poured  in  a  very  thin  section  against  a  metal  shell 
will  tend  to  be  harder  than  if  poured  in  a  thicker  section. 

Turning  again  to  Table  27,  two  striking  features  are:  first,  the  wide  varia- 
tions of  the  hardness  numbers,  and  second,  the  increase  in  hardness  after  a 
compress  in  of  1/8  inch  of  the  specimen.  White  metals  run  from  18  to  about 
35  Brinell  hardness  number,  the  highest  being  for  Souther  babbitt.  The 
bronzes  run  from  40  to  106  Brinell  hardness  number,  the  latter  being  for  a 
maganese  bronze.  This  seems  to  be  an  unusually  high  number  for  a  second 
determination  only  gives  68.  This  latter  is  closely  approached  by  two  of  the 
Cramp's  mixtures  and  exceeded  by  two  other  bronzes. 

Plumbic  bronze,  which  is  a  half-and-half  mixture  of  lead  and  copper,  has 
a  Brinell  hardness  number  of  21.8,  thus  being  in  the  range  of  the  babbitts. 


MATERIALS  FOR  BEARINGS 


53 


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54 


BEARINGS  AND  THEIR  LUBRICATION 


Composition  in  per  cent. 

Copper 

Lead 

Tin 

Antimony 

Nickel 

Phosphorus 

Sulphur 

Zinc 

64 

79.7 
64.75 

SO 

2.25 

60.67 

3-7 

2.5 

7 
5.55 

30 
95 
30 
SO 

O.IS 
32.97 

0.25 

5 

10 

5 

I 

Phosphor  bronze 

Cyprus  bronze 

0.8 

0.2s 

Parsons  white  brass 

64.9 
4.6 
88.89 
58.38 
84 
88.33 

32.9 

Demo  bronze 

2. 1 

Standard  babbitt 

7.41 

Shonberg  M.  M.  metal.  . . 
Souther  babbitt 

38.93 

9 
II .  II 



Table  28. — Analysis  of  Bearing  Metals  {Pennsylvania  Railroad  Laboratory). 

The  variation  in  babbitt  metals  purchased  on  brand  only  has  been  pointed 
out  on  page  46.  Thus  a  careful  purchaser  will  buy  on  specification,  not  on 
brand  merely,  and  at  the  same  time  will  watch  the  phyvsical  properties  of  the 
metals  furnished. 

A  similar  course  will  be  followed  by  all  who  are  careful  about  the  bronzes 
used  for  bearings. 


Name  of  metal 


Composition  in  per 

cent. 

Copper 

Tin 

Lead 

■ 

Phos- 
phorus 

Arsenic 

Phosphor  bronze '     79  . 7  10 

Copper  tin  (ist  experiment) ;     87  . 5  12.5 

Copper  tin  (2d  experiment) same 

Copper  tin  (3d  experiment . ) same 

Arsenic  bronze  (ist  experiment).  .      89.2  10 

Arsenic  bronze  (2d  experiment) ...       79.2  10 

Arsenic  bronze  (3d  experiment). . .       79 . 7  10 

"K"  bronze  (ist  experiment).  ...       77  10.5 

"K"  bronze  (2d  experiment) same 

Alloy ''B" 77  8 


9-5  o»      

-.  .  .  .      o.o3 

7         o.  oS 

9-5    ooS 

12.5     

15 


Relative 
wear  in 
per  cent. 


100 
148 
153 
147 

142 

115 

lOI 

92 

92. 

86. 


The  physical  properties  of  the  phosphor  bronze  are  tensile  strength  30,000  pounds  per 
square  inche,  elongation  6  per  cent. 

Physical    properties  of    the  alloy  "B"  are  tensile  strength  24,000  lb.  per  square  inche, 
elongation  1 1  per  cent. 

Table  29. — Relative  Wear  of  Various  Bronzes  {Dudley). 


MATERIALS  FOR  BEARINGS 


55 


WEAR  OF  CAR  BRASSES 

Dr.  Dudley  gives  the  wear  of  car  brasses  as  a  loss  of  i  lb.  of  metal 
for  every  18,000  to  25,000  miles  run.  From  his  experiments  to  determine  the 
relative  wear  of  different  alloys  in  actual  railroad  service  Table  29  has  been 
arranged. 

Referring  to  the  last  column  it  is  seen  that  the  relative  wear  in  per  cent, 
decreased  steadily  with  an  increase  in  the  proportion  of  lead,  and  that  the  wear 
of  the  alloys  containing  no  lead  was  far  greater  in  every  case  than  those  contain- 
ing that  metal. 


Copper 

Tin 

Lead 

Wear  in  grams 

8^76 

14.  no 

0 .  2800 

90 

95 
90 

6.7 
01 

9 
4. 

45 
95 
62 

0.1768 
0.0776 
0.0542 

82 

4 

4.82 

85 

12 

4 

64 

10.64 

0.0380 

81 

27 

5 

17 

14.14 

0.0327 

68 

71 

5 

24 

26.67 

0.0204 

64 

34 

4 

70 

31.22 

0.0130 

Table  30. — Relative  Wear  of  Copper-tin  and  Copper-tin-lead  Alloys  {Clamer). 


Copper 

Tin 

Lead 

Zinc 

Wear  in  grams 

85.12 

4.64 
5-28 

4-71 
5.62 
4.68 

10.64 

10.25 
10.30 
11.42 
10. 6i 

0.0380 
0.0415 
0.0466 
0 . 047  2 
0.0846 

82.27 
79.84 
77-38 
74.28 

2.07 

5-44 

6.54 

II  .04 

Table 


:elative  Wear  of  Copper-tin-lead-zinc  Alloys  (Clamer). 


Tables  30  and  31  are  from  Clamer's  experiments.  The  first  shows  plainly 
the  decrease  of  wear  with  the  increase  of  the  proportion  of  lead,  while  the 
second  shows  the  increase  of  wear  with  the  increase  of  zinc. 

Recently  compiled  data  of  the  wear  of  tender  truck  journal  bearings  are 
given  in  Table  32. 


S6 


BEARINGS  AND  THEIR  LUBRICATION 


Kind 

Size 

in. 

Load  in  lbs. 

per  sq.  in. 

projected 

area 

Wear  per 
bearing  in 

lbs.  per 
1000  miles 

Kind  of 
tender 

Plain  or 

filled 
bearings 

Composition 

Plastic  bronze 

4    1/4x8              t^t 

0.01343 

1 

Plain                     PI  a  stir     Rrnn7.f> 

Phosphor  bronze. . 

4    1/4x8 

383 

0.01700 

Plain 

Copper 64% 

Lead 30% 

Tin 5% 

Nickel 1% 

4   1/4x8 

383 

0.01200 

Plain 

Phosphor  bronze. . 

4   1/4x8 

383 

0.01814 

riain 

Cyprus  bronze. .  .  . 

4   1/4x8 

383 

0.01600 

Plain 

Copper 79.70% 

Lead 9.50% 

Tin             .         10  00% 

Phosphor  bronze. . 

4    1/4x8 

383 

0.02043 

Plain 

Phosphor 0.80% 

Cyprus  Bronze 

Plastic  bronze. .  .  . 

5    1/2XIC 

395 

0.02159 

5500  gall. 

Plain 

Standard  filled  .  .  . 

S    1/2XIC 

395 

0.02348 

5 5 00  gall. 

Filled 

Copper 64.75% 

Lead 30.00% 

Tin 5.00% 

Sulphur 0.  25% 

Standard  filled .  .  . 

S    1/2XIC 

395 

0. 02431 

5 5 00  gall. 

Filled 

Lightened  filled. .  . 

5     T/2XTC 

395 

0.02437 

5500  gall. 

Filled 

Standard  filled .  .  . 

5     I/2XIC 

420 

0.03542 

7000  gall. 

Filled 

Filling  Metal 

Lead 87% 

Antimony 13% 

Lightened  fillod. .  . 

S    1/2x10 

429 

0.03572 

7000  gall. 

Filled 

4  1/2x8  inches  used  in  both  passenger  and  freight  service. 

5  1/2x10  inches  5500  gallon  tenders — passenger  service. 
5  1/2x10  inches  7000  gallon  tenders — freight  service. 

Table  32. — Wear  of  Tender  Truck  Journal  Bearings 
(Pennsylvania  Railroad  Laboratory). 


REPRESENTATIVE  PRACTICE  IN  THE  USE  OF  BEARING  METALS 

The  bearing  metal  known  as  the  standard  of  the  Bureau  of  Steam  Engineer- 
ing of  the  United  States  Navy,  also  called  "anti-friction"  or  "anti- attrition" 
metal,  has  this  composition: 

Best  refined  copper 3.7  per  cent. 

Banca  tin 88 . 8  per  cent. 

Regulus  of  antimony 7.5  per  cent. 

The  percentages  are  by  weight.  The  mixture  must  be  well  fluxed  with 
borax  and  rosin  in  mixing. 

The  mixing  of  this  anti-friction  metal  is  a  trick  which  must  be  learned. 
The  best  practice  is  to  melt  the  copper,  tin  and  antimony  separately,  adding 
the  tin  to  the  copper  and  the  antimony  to  this  mixture,  fluxing  it  with  borax 
with  the  proportion  of  about  i  1/2  lb.  to  175  lb.  of  the  mixture;  but  satisfactory 


MATERIALS  FOR  BEARINGS 


57 


results  are  obtained  by  melting  the  copper  first,  dropping  the  cold  tin  into  the 
melted  copper  and  adding  the  antimony,  which  has  been  separately  melted. 
This  metal  is  carefully  skimmed  before  pouring,  and  is  poured  into  pigs  and 
carried  into  stock  as  it  stands. 

The  journal  bronze  used  on  battleships  of  the  United  States  Navy  has  this 
composition:  Copper  82  to  84,  tin  12.5  to  14.5,  zinc  2.5  to  4.5,  iron  (max.) 
0.06,  lead  (max.)  i.oo,  all  in  per  cent.,  with  a  normal  of  83-13  1/2-3  1/2.  It 
is  used  for  bearings,  bushings,  sleeves,  slides,  guide  gibs,  wedges  on  water- 
tight doors  and  all  parts  subject  to  considerable  wear. 

The  anti-friction  metals  commonly  specified  by  the  United  States  War 
Department  are  Parson's  White  Brass,  Genuine  Babbitt,  Magnolia  Metal, 
Phoenix  Metal,  and  Shonberg's  "M.M."  Parson's  White  Brass  from  one 
analysis  is  composed  of:  tin  64.90,  lead  0.15,  copper  2.25,  zinc  32.93;  Shonberg 
M.M.  white  bronze:  tin  38.38,  zinc  38.93,  copper  2.5,  lead  0.25. 

This  comparison  of  the  physical  properties  of  three  of  these  metals  is  taken 
from  the  Journal  of  the  American  Society  of  the  Naval  Engineers,  Volume  XIX. 


Naval  white  brass 


Parson's  white  brass 


Magnolia  metal 


Original      diameter,      inches 
height,  inches. 

Elastic  limit  (about) ,  lbs 

Final  compression  loads,  lbs. . 
Final  height,  inches 


0.7S 

0.75 

0.7S9 

0.758 

0.751 

8,000 

8,000 

5.000 

5,000 

5,000 

36,000 

37.000 

35. 000 

36,800 

37.000 

0.2 

0.192 

0.180S 

0..8 

0.  i6r 

0.748 

4,500 
36,500 

0.166 


The  Westinghouse  Electric  &  Manufacturing  Company  has  developed  most 
careful  practice  in  preparing  and  handling  bearing  metals.  In  general,  two 
kinds  of  babbitt  are  used ;  one  having  a  tin  base  and  the  other  a  lead  base.  Not 
only  are  the  proportions  of  the  alloys  carefully  watched,  but  also  the  physical 
properties. 

A  standard  test  block  to  determine  hardness  is  in  the  form  of  a  disk  2-1/4 
in.  in  diameter,  3/4  in.  thick,  and  surf aced .  both  sides.  When  tested  by 
the  Brinell  method,  the  genuine  babbitt  shows  a  hardness  number  of  about  25; 
similarly,  the  softer  gives  a  number  of  about  21. 

Care  is  used  not  to  overheat  the  metals,  for  the  harder  the  pouring  tempera- 
ture should  not  exceed  890°  F.;  the  corresponding  minimum  temperature  is 
about  660°  F.  In  practice  an  effort  is  made  to  keep  the  pouring  temperature 
about  midway  of  there  limits.  For  the  softer  metal  the  corresponding  temper- 
atures are  1000  and  840°  F.  This  alloy  is  more  easily  abused  by  overheating 
than  the  harder  one. 

In  general  the  softer  metal  is  used  for  the  regular  run  of  work  and  the  harder 
in  railway  motor  bearings  and  bearings  subjected  to  pound.     Experience  has 


58  BEARINGS  AND  THEIR  LUBRICATION 

shown  that  for  conditions  of  nearly  steady  pressure  the  softer  metal  is  in  many 
instances  preferable  to  the  harder  one,  being  less  liable  to  overheating  in  service. 
Albert  E.  Guy  gives  the  composition  of  the  high-speed  babbitt  used  in 
De  Laval  steam  turbines  as:  copper  lo,  tin  80,  and  antimony  10  per  cent. 
For  low  speeds  the  metal  used  is:  lead  77,  tin  6  and  antimony  17  per  cent. 

RAILROAD    PRACTICE 

Representative  practice  of  the  Pennsylvania  Railroad  Company  is  as  follows: 
For  all  axle  bearings  of  passenger  and  freight  cars,  antimonial  lead  lining  on 
ExB  bronze  shells;  for  locomotive  driving  wheel  axle  bearings,  crank  pin  bear- 
ings, cross-head  pin  bearings,  and  small  bearings  generally,  ExB,  phosphor-, 
or  plastic  bronze;  for  cross-head  shoes,  block  tin;  for  four-wheel  engine  truck 
boxes,  standard  babbitt  strips  set  in  bronze. 

The  specifications  for  journal  bearing  lining  metal  calls  for  a  homogeneous 
alloy  of  lead  and  anitmony  as  free  as  possible  from  every  other  substance,  and  of 
the  following  composition:  Lead  87  per  cent.,  antimony  13  per  cent.  The 
rejection  limits  by  analysis  are:  Lead  less  than  86  and  more  than  88  per  cent., 
tin  and  antimony  less  than  12  and  more  than  14  per  cent.,  tin  more  than  2 
per  cent.,  copper  more  than  0.5  per  cent.,  or  if  the  sum  of  the  amounts  of  lead, 
tin,  antimony  and  copper  is  less  than  99.5  per  cent. 

The  specifications  for  phosphor-bronze  bearing  metal  call  for  a  homo- 
geneous alloy  of  copper,  tin,  lead,  phosphorus,  as  free  as  possible  from  every 
other  substance,  and  of  the  following  composition:  Copper  79.7  per  cent.,  tin 
10  per  cent.,  lead  9.5  per  cent.,  phosphorus  0.8  per  cent.  The  rejection  limits 
are  as  follows:  Lots  will  not  be  accepted  if  the  analysis  as  above  described  gives 
results  outside  the  following  limits:  Tin,  below  9  per  cent,  or  over  11  per  cent.; 
lead,  below  8  per  cent,  or  over  11  per  cent.;  phosphorus,  below  0.7  or  over  i  per 
cent. ;  nor  if  the  metal  contains  a  sum-total  of  any  other  substances  than  copper, 
tin,  lead  and  phosphorus  in  greater  quantity  than  i  per  cent. 

The  specification  for  standard  babbitt  calls  for  a  homogeneous  alloy  of  tin, 
antimony  and  copper,  having  the  following  composition  in  per  cent.:  tin 
88.89,  antimony  7.41,  copper  3.7. 

The  ExB  bronze  analysis  is: 

Copper  76.75  per  cent.,  limits 75  -75  to  77  .75 

Lead  15  per  cent.,  limits 13.5    to  16 . 5 

Tin  S  per  cent.,  limits 7         to    9 

Phosphorus  0.25  per  cent.,  limit .' not  below  0.20 

Impurities,  limit not  above    0.75 

Robert  Garbe,  in  Die  Danpflokomotiven  der  Gegenwart,  page  345,  gives  the 
composition  of  a  white  bearing  metal  that  has  been  used  successfully  for  50 
years  by  the  Prussian  State  Railways;  during  that  time  it  has  remained  un- 


MATERIALS  FOR  BEARINGS  "  59 

changed.  It  is  used  not  only  for  the  axle  bearings  of  locomotives,  but  also 
in  both  freight  and  passenger  cars. 

It  is  made  of  copper,  antimony  and  tin,  in  the  following  manner:  i  kg. 
(2.2  lb.)  of  copper  is  melted  together  with  2  kg.  (4.4  lb.)  antimony  (regulus) 
and  6  kg.  (13.2  lb.)  of  Strait's  tin.  The  antimony  is  added  when  the  copper  is 
melted,  and  when  both  metals  are  fluid,  the  tin.  This  alloy  is  cast  in  thin 
plates  and  each  9  kg.  (19.8  lb.)  is  remelted  with  9  kg.  (19.8  lb.)  of  pure  tin. 
The  whole  is  then  cast  in  plates  15  mm.  (0.6  in.)  thick  and  is  therewith  ready 
for  use.  Larger  quantities  than  those  mentioned  above  ought  not  to  be  melted 
together  at  one  time. 

The  individual  constituents  of  the  alloy  must  be  as  pure  as  possible.  Thus 
the  antimony  should  contain  at  the  most  i  per  cent,  of  impurities  and  among 
them  not  more  than  o.i  per  cent,  of  arsenic.  Tin,  on  the  other  hand,  should 
contain  not  more  than  0.2  per.  cent,  of  impurities.  Furthermore,  especially, 
lead  and  zinc  ought  not  to  be  mixed  in. 

The  constituents  stated  in  per  cent,  are:  tin  83.33,  copper  5.55,  antimony 
II. II. 

G.  H.  Clamer  sums  up  American  railroad  practice  in  1907  in  the  Proc.  A. 
S.  T.  M.,  Vol.  VII,  page  302.  Continued  experiments  have  shown  that  tin 
could  be  reduced  and  lead  increased  beyond  the  amounts  given  by  Dr.  Dudley, 
and  that  a  satisfactory  bearing  metal  could  be  made  composed  of  65  per  cent, 
copper,  5  per  cent,  tin  and  30  per  cent.  lead.  This  alloy  is  largely  sold  under 
the  name  of  'Aplastic  bronze."  It  has  a  compressing  strength  of  about  15,000 
lb.  to  the  square  inch  and  is  found  to  operate  without  distortion  in  the  bearings 
of  the  heaviest  locomotives  for  the  driving  brasses,  rod  brasses  and  bushings, 
and  for  bearings  of  cars  of  100,000  lb.  capacity.  Railroad  specifications 
cover  bearing  alloys  having  a  tin  content  of  from  8  to  10  per  cent,  and  lead  from 
10  to  15  per  cent. 

As  railroad  practice  yields  a  large  amount  of  old  material,  or  scrap,  which 
must  be  used,  many  car  brasses  are  made  from  old  metal.  These  contain 
copper  65  to  76  per  cent.,  tin  2  to  8  per  cent.,  lead  10  to  18  per  cent.,  and  zinc 
from  5  to  20  per  cent.  It  is  estimated  that  such  bearings  constitute  from  50  to 
75  per  cent,  of  all  the  car  bearings  in  use. 

The  Mesta  Machine  Company  on  rolling  mill  work  uses  two  grades  of 
babbitt  and  a  bronze.  For  the  general  run  of  work,  a  lead  babbitt  is  satis- 
factory having  this  composition  in  per  cent.,  lead  75,  tin  12.5,  antimony  1^.$. 
For  high  rubbing  speeds  a  mixture  is  made  of  i  part  of  the  above  and  2  parts 
of  genuine  babbitt.  This  genuine  babbitt,  alone,  is  used  on  rolling  mill  engines 
and  in  bearings  subjected  to  shock  and  pound.  Its  composition  in  per  cent,  is: 
tin  82,  copper  5.4,  antimony  12.6.  The  bronze  is  a  tough  copper-tin-lead 
alloy  very  similar  to  Pennsylvania  Railroad  metal. 


6o  BEARINGS  AND  THEIR  LUBRICATION 

In  the  rolling  mills  for  repairs  any  babbitt  is  considered  satisfactory  having 
about  lo  per  cent,  tin,  12  to  14  per  cent,  antimony,  and  the  rest  unspecified. 
"  Copper  alloy"  is  also  used  for  roll  bearings. 

AUTOMOBILE    PRACTICE 

The  alloys  division  of  the  Standards  Committee  of  the  Society  of  Auto- 
mobile Engineers  in  their  report  for  June,  191 1,  specifies  three  bearing  metals 
as  follows: 

Babbit  Metal,  Specification  No.  24 

Tin 84  per  cent. 

Antimony 9  per  cent. 

Copper 7  per  cent. 

A  variation  of  i  per  cent,  either  way  will  be  permissible  in  the  tin,  and  0.5 
per  cent,  either  way  will  be  permissible  in  the  antim.ony  and  copper.  The  use  of 
other  than  virgin  metals  is  prohibited.  No  impurity  will  be  permitted  other 
than  lead,  and  that  not  in  excess  of  0.25  per  cent. 

Note:  This  grade  of  babbitt  is  special,  owing  to  the  large  amount  of  copper 
contained  therein.  It  is  used  for  the  connecting-rod  bearings  of  gasoline 
motor  bearings,  locomotive  work,  or  for  any  service  where  machinery  designers 
are  confronted  with  severe  operating  conditions. 

White  Brass,  Specification  No.  25 

Copper 3  .  00  to    6 .  00  per  cent. 

Tin,  not  less  than 65  .00  per  cent. 

Zinc 28 . 00  to  30. 00  per  cent. 

Metal  containing  more  than  0.25  per  cent,  impurities  may  be  rejected. 
Note:  This  alloy  gives  good  results  in  automobile  engines,  but  provision 
should  be  made  to  have  it  generously  lubricated. 

Phosphor  Bronze  Bearing  Metal,  Specification  No.  26 

Copper 80 .  00  per  cent. 

Tin 1  o. GO  per  cent. 

Lead i o. 00  per  cent. 

Phosphorus o. 05  to  o. 25  per  cent. 

Impurities  in  excess  of  0.25  per  cent,  will  not  be  permitted. 

Note:  This  is  a  metal  similar  to  that  specified  by  many  railroads  for  various 
purposes.  It  is  an  excellent  composition  where  good  anti-frictional  qualities 
are  desired,  standing  up  exceedingly  well  under  heavy  loads  and  severe  usage. 
It  should  be  used  only  against  hardened  steel  in  automobile  construction. 


MATERIALS  FOR  BEARINGS 


6i 


Red  Brass,  Specification  No.  27 
Copper 85  .00  per  cent. 

Tin i 5 .  00  per  cent. 

Lead 5 .00  per  cent. 

Zinc 5 .  00  per  cent. 

A  tolerance  of  i  per  cent,  plus  or  minus  will  be  allowed  in  the  above  percent- 
ages.    Impurities  in  excess  of  0.25  per  cent,  will  not  be  permitted. 

Note:  A  high  grade  of  composition  metal,  and  an  excellent  bearing  where 
speed  and  pressure  are  not  excessive.  Largely  used  for  light  castings,  and 
possesses  good  machining  qualities. 

ALLOYS  FOR  METALLIC  PACKING  . 

In  metallic  packing  for  piston  rods,  valve  stems,  and  turbine  shafts,  we  find 
a  use  of  alloys  similar  to  their  use  as  bearing  metals  but  with  this  important 


Plan 


Shaft  Sleeve. 

Packing  against 
Pressure. 

Springs  placed 
•^ about  4"apart. 

— Packing  against 
Vacuum. 


Driving  Fit -^  |-1-^-'* 


5C 


Y2 


Metallic  Back.     f^»»^^ 
Notches  in  Back  L  _o7^'l  _J 
to  hold  Metallic   '    '  "^^  '^^  „ 
Packing.  Section  6-6 

Metallic  Packing. 
Brass  Band. 


FIG.    8.       CUNNINGHAM    METALLIC    PACKING    FOR   VTCRTICAL    STEAM   TURBINE. 


62 


BEARINGS  AND  THEIR  LUBRICATION 


difference:  In  packings  the  metal  lining  is  not  intended  to  carry  any  weight 
whatever  but  closes  tightly  around  the  shaft  or  rod  and  makes  a  steam-tight 
joint.  However,  as  the  rod  and  shaft  are  in  motion,  we  have  a  condition  of 
ordinary  sliding  friction,  but  differing  again  from  bearing  practice  in  that  the 
parts  are  heated  to  a  high  temperature  by  steam. 


FIG.    9.      SECTION    OF   LABYRINTH   PACKING    FOR   STEAM   TURBINE    SHAFT. 

Fig.  8  shows  the  Cunningham  turbine  shaft  packing  as  used  on  Curtis 
verticle  turbines  in  place  of  carbon  packing.  It  consists  of  cast-iron  quadrants 
filled  with  Allan  ''red  metal,"  50  per  cent,  lead  and  50  per  cent,  copper,  fitted 
together  and  surrounded  by  a  brass  band  with  helical  tension  springs. 

The  General  Electric  Company  uses  a  similar  packing  but  arranged  in  the 
form  of  a  labyrinth,  as  shown  in  Fig.  9.  Both  of  these  packings  are  run  without 
lubrication  and  are  peculiarly  well  adapted  for  high-speeds  and  high  temper- 
atures. 


Babbitt  Rings  in  Halves  '/is  Cut-out 

Br  133 
Dowel  Pin^  \'  f^^^^^^m^''^^ 


Babbitt  Rings  in  Halves 

Brass 


II— 4-4-J-^^ 

Multiangular 
PaclsmS—. 


■^"^Single^  Angle     f 
_  _  JPackiQg' I 


FIG.    10.      ARRANGEMENT    OF    KING    METALLIC    PACKING    FOR    LOCOMOTIVE    PISTON    RODS. 


To  show  the  practice  in  regard  to  piston  rods  and  valve  stem  packings, 
Fig.  10  is  given.  These  are  the  "single  angle"  and  "multi- angular"  packing 
of  the  United  States  Metallic  Packing  Company,  as  applied  in  locomotive 
service.  The  babbitt  mixture  used  for  the  single  angle  packing  is  100  parts 
tin,  9  parts  copper,  and  6  parts  antimony.     For  the  multi-angular  packing  the 


MATERIALS  FOR  BEARINGS  63 

mixture  is  83  1/3  parts  lead,  8  1/3  parts  tin,  and  8  1/3  parts  antimony.     Allan 
"red  metal"  is  also  used. 

In  addition  to  these  solid  metal  packings,  there  are  many  soft  packings 
composed  of  metallic  wool  or  fibers  pressed  together  into  a  cord  or  rope  and 
used  in  place  of  ordinary  fiber  packings.  They  are  composed  largely  of  lead 
with  the  addition  of  a  small  amount  of  tin  or  copper.  A  representative  analysis 
is  88  lead  and  12  copper. 

VARIATIONS  IN  BEARING  METALS 

In  the  discussion  of  babbitt  metals  it  is  pointed  out  that  many  such  "bab- 
bitts," placed  on  the  market  in  fancy  shaped  ingots  and  under  attractive 
names,  are  nothing  but  waste  metals  from  type  foundries  and  "clean  up" 
from  smelters.  Of  course,  such  metals  vary  in  composition.  If  a  particular 
mixture  is  desired  for  bearing  purposes,  it  should  be  purchased  from  a  reputable 
dealer  according  to  a  fixed  analysis. 

In  the  use  of  bronzes  there  is  not  only  the  variations  in  proportions  of  the 
constituent  elements  to  be  taken  into  consideration,  but  also  the  way  in  which 
the  mixture  is  compounded.  Under  different  foundry  conditions  the  same 
constituent  metals  in  identical  proportions  may  give  alloys  with  different 
physical  properties.  In  adopting  a  bronze  bearing  metal  one  of  the  important 
points  is  to  select  an  alloy  that  is  easily  handled  under  ordinary  foundry  con- 
ditions. If  a  precise  mixture  is  desired,  reputable  metal  dealers  and  foundry- 
men  can  be  found  who  will  guarantee  to  furnish  either  ingot  metal  or  castings, 
in  which  none  of  the  constituent  metals  will  vary  more  than  one  half  of  i  per 
cent,  from  the  specification. 

WOOD  AS  A  BEARING  MATERIAL 

But  little  need  be  said  in  regard  to  the  use  of  wood  for  bearings.  In  the 
early  days  of  mill  construction,  lignum  vitai  was  successfully  used  for  the 
bearings  of  mill  shafting.  This  wood  has  a  peculiar  greasy  character  which, 
coupled  with  its  extreme  hardness  and  density,  rendered  it  successful  for 
moderate  pressure  and  comparatively  low  speed. 

Two  other  uses  are  deserving  of  mention:  One  is  in  the  step  bearings  of 
water  turbines  where  the  step  is  constantly  flooded  with  water.  If  the  pres- 
sures and  speed  were  not  too  high  it  has  proved  reasonably  successful  in  this 
service. 

The  second,  and  at  the  present  time  practically  its  only  application,  is  in  the 
tail  shaft  bearings  of  steam  ships  and  submerged  guide  bearings  of  water  wheels. 
Here  it  is  commonly  put  in  in  the  form  of  staves,  and  as  the  bearing  is  open  to 
the  water  it  is  constantly  lubricated  and  cooled.     Other  kinds  of  wood  have 


64  BEARINGS  AND  THEIR  LUBRICATION 

been  used  to  a  limited  extent,  such  as  oak,  wild  pear,  and  apple  tree  wood. 
Several  attempts  have  been  made  to  put  on  the  market  wooden  bushings 
soaked  with  a  wax  or  grease  to  render  it  unnecessary  to  oil  them.  They  are 
usually  described  as  ''oilless  bearings."  For  very  light  pressures  they  have 
had  a  certain  amount  of  success,  although  their  use  is  by  no  means  general. 

GRA.PHITE   AS  A   BEARING   MATERIAL 

Graphite  bearings  usually  consist  of  a  brass  or  bronze  shell  having  a  series 
of  openings  or  slots  into  which  graphite  in  the  form  of  a  paste  is  forced  under 
pressure  and  allowed  to  harden.  Such  bearings  are  intended  to  be  run  without 
oil,  and  are  especially  adapted  for  positions  where  oiling  would  be  very  difficult 
or  under  conditions  of  considerable  heat.  In  operation  the  bronze  shell 
supports  the  shaft  and  the  graphite  acts  as  a  lubricant. 

Another  form  of  graphite  bearing  consists  in  casting  a  piece  of  netting 
having  a  number  of  slugs  of  graphite  attached  thereto  into  the  bearing  lining 
if  of  babbitt,  or  into  the  box  or  brass  if  of  bronze.  This  type  is  not  intended 
to  run  without  lubrication,  but  the  addition  of  the  graphite  in  the  form  of 
these  slugs  equally  spaced  throughout  the  bearing  surface  is  intended  to  reduce 
friction  by  adding  a  valuable  solid  lubricant. 

MISCELLANEOUS   MATERIALS 

Hard  fiber  is  used  for  small  thrust  collars  and  with  low  pressures  and 
ample  lubrication  is  satisfactory. 

Knife-edge  V's  in  weighing  machines  and  computing  scales  are  frequently 
made  of  agate  with  carefully  polished  surfaces. 

The  most  frictionless  bearings  that  are  produced — if  bearings  they  can  be 
called — are  mercury  baths.  They  have  been  used  in  French  lighthouses  and 
to  support  certain  parts  of  astronomical  instruments.  They  permit  of  the 
freest  motion  of  the  floating  members.  This  is  an  example  of  perfect  fluid 
support.  As  a  metal  that  will  not  amalgamate  must  be  used  for  the  float  and 
casing,  these  parts  are  made  of  cast-iron.  The  specific  gravity  of  mercury  is 
13.58. 


SECTION  IV 

ALLOWABLE  BEARING  PRESSURES,  SPEEDS  AND 
TEMPERATURES 

In  Section  II  it  is  pointed  out  that  values  of  the  coefficient  of  friction  for 
types  of  bearings  and  various  kinds  of  lubrication  are  none  too  accurately  known. 
The  same  condition  exists  to  a  certain  extent  for  allowable  bearing  pressures. 
However,  good  practice  is  given  in  Table  33  for  several  kinds  of  machines 
and  service.  For  light  machinery  the  unit  pressures  are  so  small  as  to  be 
negligible. 


Kind  of  bearing  and  condition  of  operation 


Allowable  bearing  pressure  in  pounds 
per  square  inch  of  projected  area 


Bearings  for  very  low  speeds  and  intermittent  service 
as  in  turntables  and  bridges. 


7000  to  9000 


Railroad  Practice 


Locomotive  cro?s-head  pin  bearings 

Locomotive  crank  pin  bearings 

Locomotive  driving  wheel  journal  bearings . 

Car  axles  bearings 

Tender  axle  bearings 


3000 

to 

4000 

1500 

to 

1700 

to 

550^ 

300 

to 

325 

to 

425 

British  Railway  Practice 


Locomotive  crank  pin  bearings 

Locomotive  cross-head  pin  bearings. 
Locomotive  driving  axle  bearings.  .  . 
Car  axle  bearings 


...  to  1400 
...  to  2000 
250  to  300 
...  to    330 


Table  32. — Allowable  Bearing  Pressures  for  Machinery  Bearings. 

1  Pennsylvania  Railroad  Class  E6  locomotives  have  a  bearing  pressure  on  bearings  of  main  drivers 
of  556  lb.  per  square  inch. 

S  6s 


66 


BEARINGS  AND  THEIR  LUBTICATION 


Eind  of  bearing  and  condition  of  operation 


Allowable  bearing  pressure  in  pounds 
per  square  inch  of  projected  area 


United  States  Naval  Practice 


Main  engine  bearings 

Main  engine  crank  pin  bearings .  . 

Steam  turbine  bearings 

Thrust  bearings  for  torpedo  boats. 


275  to 

400 

400  to 

500 

...to 

85 

.  .  to 

50 

For  weight  alone. 


Merchant  Marine  Practice 


Main  engine  bearings 400  to    500 

Main  engine  crank  pin  bearings ]       400  to    500 


High-speed  Stationary  Engine  Practice 


Main  bearings , 

Main  bearings 

Crank  pin  bearings,  overhung  cranl. 
Crank  pin  bearings,  center  crank. . 
Cross-head  pin  bearings , 


60  to    120      For  dead  load. 


150  to    250 

900  to  1500 

400  to    600 

1000  to  1800 


For  steam  load. 


Slow  Speed  Stationary  Engine  Practice 


Main  bearings 

Main  bearings 

Crank  pin  bearings 

Cross-head  pin  bearings. 


For  dead  load. 
For  steam  load. 


Air  Compressor  Practice  ^ 


Straight  line,  steam  driven,  100  lb.  steam  and  air. 


Main  bearings 

Crank  pin  bearings 

Cross-head  pin  bearings. 


160  to  237 
565  to  700 
628  to  820 


Table  33. — Continued. — Allowable  Bearing  Pressures  for  Machinery  Bearings. 
1  Canadian  Rand  Company. 


ALLOWABLE  PRESSURES 


67 


Kind  of  bearing  and  condition  of  operation 


Allowable  bearing  pressure  n  pounds 
per  square  inch  of  projected  area 


Straight  line,  belt  driven,  center  crank,  100  lb.  steam  and  air. 


Main  bearings !       122  to  220 

Crank  pin  bearings I       244  to  402 

Cross-head  pin  bearings |      400  to  785 


Straight  line,  belt  driven,  side  crank,  100  lb.  steam  and  air. 


Main  bearings |       178  to  227 

Crank  pin  bearings .  .  .  .  |       628  to  825 

Cross-head  pin  bearings [      628  to  825 


Straight  line,  steam  driven,  side  crank,  100  lb.  steam  and  air. 


Main  bearings 198  to  227 

Crank  pin  bearings 462  to  825 

Cross-head  pin  bearings 462  to  825 

i 


Duplex,  Meyer  cut-off,  steam-driven,  100  lb.  stearr  and  air. 


Main  bearings 

Crank  pin  bearings 

Cross-head  pin  bearings. 


157  to  200 
644  to  855 
850  to  1370 


Duplex  Corliss  valve  gear,  steam  driven,  100  lb.  steam  and  air. 


Main  bearings 

Crank  pin  bearings 

Cross-head  pin  bearings 

Direct-connected,  motor  driven  main  bearings 


Gas  Engine  Practice 


Main  bearings 

Crank  pin  bearings 

Cross-head  pin  bearings. 


500  to  700 
1500  to  1800 
1500  to  2000 


Table  s^. — Cw^/inwei.— Allowable  Bearing  Pressures  for  Machinery  Bearings. 


68 


BEARINGS  AND  THEIR  LUBRICATION 


Kind  of  bearing  and  condition  of  operation 


Allowable  bearing  pressure  in  pounds 
per  square  inch  of  projected  area 


Electrical  Machinery  Practice 


Generator  and  motor  bearings 

Main  engine  bearings,  driving  generators 

Horizontal  steam  turbine  bearings 

Vertical  steam  turbine  steps 


30 

to 

80 

40 

to 

80 

400  to 

60 

00 

to  1000 

Rolling  Mill  Practice  ^ 


Rubbing  velocity  in 
feet  per  minute 


Pinion  housing  bearings I  30  to  50  ^ 

Roll  housing  beanngs i  100  to  2000  ^ 

Table  roller  bearings j  30  to  50 

Table  line-shaft  bearings 1  30  to  50 

Main  bearings  of  shears 1  1800  to  2500 


350  to  600 
350  to  600 

150 

150 
50  to    65 


Miscellaneous  Practice 


Bearings  for  slow  speed  and  intermittent  load  as  in 
punch  presses,  shears,  and  the  like. 

Main  bearings  of  slow  speed  pumping  engines 

Heavy  line-shaft  bearings,  bronze  or  babbitt  lined. . 

Light  line-shaft  bearings,  cast-iron 

Heavy  slow  speed  step  bearings 

Drill  press  thrust  collars 

Angular-thrust  bearing  tor  boring  mill  tables 


3000 

to 

4000 

to 

600 

100 

to 

150 

15 

to 

25 

to 

2000 

to 

325 

— 

to 

75^ 

Table  33. — Continued. — ^Allowable  Bearing  Pressures  for  Machinery  Bearings. 

Tables  33A  to  33F,  inclusive,  are  from  an  article  by  G.  W.  Lewis  and 
A.  G.  Kessler  published  the  American  Machinist.  They  give  main  bearing 
crank  pin  bearing,  and  wrist  or  piston  pin  bearing  pressures  for  large  stationary 
gas  engines,  both  horizontal  and  vertical.     Four  maximum  explosion  pressures 

1  Mesta  Machine  Company,  Pittsburg,  Pa. 

^  These  factors  are  of  value  as  showing  good  practice,  not  for  purposes  of  design.  The  diameters 
and  lengths  of  the  bearings  are  determined  by  the  requirement  of  strength  in  the  pinion  and  roll  necks 
and  their  housings. 

^  Practice  of  BuUard  Machine  Tool  Company. 


ALLOWABLE  PRESSURES 


69 


in  pounds  per  square  inch  of  piston  face  are  assumed,  namely,  250,  300,  350, 
and  400  lb.  These  tables  give  a  portion  of  the  results  of  an  extensive  inves- 
tigation of  the  designs  of  a  large  number  of  engines  and  thoroughly  reflect 
modern  practice. 

D  =  cylinder  diameter  in  inches. 

Pm  =  maximum  explosion  pressure  in  pounds  per  square  inch  of 

piston  face. 

Dcp  =  bearing  diameter  of  crank  pin  in  inches. 

Lcp  =  bearing  length  of  crarik  pin  in  inches. 

Kcp  =  maximum  unit  bearing  pressure  in  pounds  per  square  inch. 

Acp  =  DcpX  Lcp  in  square  inches. 

Max.  Kcp  =  71     Pm  -^  (Dcp  X  Lcp)      (A) 
4 


D 

4 

8 

12 

16 

20 

Assumed 

Dcp 
Lcp 
Acp 

I   1/2 
I  5/8 
2.44 

3  1/8 
3  1/4 
10.15 

4  3/4 
4  7/8 
.3- 

63/8 
6  9/16 

41.75 

8  1/8 
8  3/8 
68 

DcpX  Lcp  =  Acp 

Pm 


Assumed 


Kcp 


1290 


1240  1220 


1150 


From  equation  A 


Pm 


300 


Assumed 


Kcp 


1550 


1485 


1450 


1450       ,       1390         From  equation  A 


Pm 


Kcp 


350 


1800       j       1730 


1710 


Assumed 


1690  1620       !  From  equation  A 


Pm 


400 


Assumed 


Kcp  2060  1980 

I  ■       i 


1950 


1930  1850       I  From  equation  A 


Table  33A. — Crank-pins  Bearing  Pressures  for  Horizontal  Stationary  Gas  Engines. 


70 


BEARINGS  AND  THEIR  LUBRICATION 


D  =  cylinder  diameter  in  inches. 

Pm  =  maximum  explosion  pressure  in  pounds  per  square  inch  of 

piston  face. 

Dcp  =  bearing  diameter  of  crank  pin  in  inches. 

Lcp  =  bearing  length  of  crank  pin  in  inches. 

Kcp  =  maximum  unit  bearing  pressure  in  pounds  per  square  inch. 

Acp  =DcpX  Lcp  in  square  inches. 


Pm 


D 

4 

8 

12 

16 

20 

Assumed 

Dcp 
Lcp 
Acp 

I  5/8 
I  5/8 
2.64 

3  1/4 
3  5/8 
II. 8 

4  7/8 

5  5/8 
27.8 

6  1/2 

7  5/8 
49-75 

8  3/16 

9  5/8 

78.75 

Dcp  X  Lcp 

250 


Assumed 


Kcp 

1190 

1065 

1035 

1015 

995 

Pm 

30c 

) 

Assumed 

Kcp 

1430 

1280 

1240 

1215 

1200 

Pm 

35c 

) 

Assumed 

Kcp 

1660 

1490 

1440 

1420 

1400 

Pm 

40c 

Assumed 

Kcp 

1920 

1720 

I  60 

1620 

1600 

Table  2>:i  B. — Crank-pin  Bearing  Pressures  for  Vertical,  Stationary  Gas  Engines. 


ALLOWABLE  PRESSURES 


n 


D       =  cylinder  diameter  in  inches. 

Dmh  =  main  bearing  diameter  in  inches. 

Lmb   =  main  bearing  length  in  inches. 

Amb  =  projected  area  main  bearing  (one)  =  DmhX Lmb. 

Pm    =  maximum  explosion  pressure  in  pounds  per  square  inch  of 

piston  face. 
Kmb  =  maximum  unit  bearing  pressure  in  pounds  per  square  inch 
considering  explosion  to  occur  on  dead  center. 


D 


i6 


20  Assumed 


Dmb 
Lmb 
Amb 


1-3 

31 

4.8s 

2.6 

6.75 

10.8 

3-4 

20.9 

52.5 

6.6 

14.9 
98.5 


8.4 

19. 1 

160.5 


Dmb X Lmb 


Vm 

25c 

> 

Assumed 

Kmb 

462 

300 

270 

25s 

244 

Pm 

30c 

Assumed 

Kmb 

553 

360 

324 

307 

293 

Pm 

350 

. 

Assumed 

Kmb 

"' 

420 

377 

358 

342 

Pm 

40c 

) 

Assumed 

Kmb 

738 

503 

430 

408 

391 

Table  33C. — Main  Bearing  Pressures  for  Horizontal,  Stationary  Gas  Engines. 


72 


BEARINGS  AND  THEIR  LUBRICATION 


D       =  cylinder  diameter  in  inches 
Dmh  =  bearing  diameter  of  main  bearing  in  inches 
Lmh   =  bearing  length  of  main  bearing  in  inches 
Amh  =  DmbxLmb 

Pm     =  maximum  explosion  pressure  in  pounds  per  square  inch  of 
piston  face. 


Pm 


Kmb 


300 


D 

4 

8 

12 

16 

20 

Assumed 

Dmb 
Lmb 
Amb 

I  1/2 
3  1/2 
5-25 

3  1/2 
63/4 
23.6 

5  1/2 
10 

55 

7  1/2 
13 
97-5 

9  1/2 
16 
152 

DmbxLmb 

267 


250 

258 


Assumed 


258 


258 


Pm 


Kmb 


359 


320 


300 


Assumed 


310 


310 


310 


Pm 


Kmb 


415 


373 


350 


Assumed 


360         I       360         !       360 


Pm 


400 


Assumed 


Kmb    I       481 


414 


414  414         t       414 


Table  33D. — Main  Bearing  Pressures  for  Vertical  Stationary  Gas  Engines. 


ALLOWABLE  PRESSURES 


73 


D       =  cylinder  diameter  in  inches 

Dwp  =  bearing  diameter  of  piston  pin  in  inches 

Lwp    =  bearing  length  of  piston  pin  in  inches 

Awp  =  projected  area  piston  pin  in  square  inches 

Pm    =  maximum  unit  explosion  pressure 

Kwp  =  maximum  unit  bearing  pressure  in  pounds  per  square  inch 

JDwp  =0.0143  D"+o.  f  (A) 

Lwp    =i.'j^Dwp  (B) 


D 

4 

8 

12 

16 

20 

Dwp 

•93 

1.62 

2.76 

4.36 

6.42 

Lwp 

1.6 

2.8 

4-77 

7-52 

II. 15       I 

Awp 

1.49 

4-54 

13.2 

32.8 

71-5 

From  Equation  (B) 
Dwp  X  Lwp 


Pm 

250 

Assumed 

Kwp 

2100 

2760 

2145 

1530 

1 100 

Pm 

300 

Assumed 

Kwp 

2530 

3320 

2570 

1840 

1320 

Pm 

35° 

Assumed 

Kwp 

2950 

3880 

3000 

2150 

1540 

Pm 

400 

Assumed 

Kwp 

3370 

4425 

3430 

2455 

1760 

Table  33E. — ^\Vrist  or  Piston  Pin  Bearing  Pressures  for  Horizontal,  Stationary 

Gas  Engines. 


74 


BEARINGS  AND  THEIR  LUBRICATION 


D       =  cylinder  diameter  in  inches 

Dwp  =  bearing  diameter  of  piston  pin 

Lwp    =  bearing  length  of  piston  pin 

Awp  =  projected  area  piston  pin  in  square  inches 

Kwp  =  maximum  unit  bearing  pressure  in  pounds  per  square  inch 

Pm    =  Maximum  unit  explosion  pressure. 

Dwp  =0.00795  ^'+  I  3/8''  (C) 

Lwp    =1.82  Dwp   (D) 


D 


Dwp 
Lwp 
Awp 


15/8 
27/8 
4.67 


16 


I  7/8  I  2  1/2  j  3  3/8 
3  3/8  I  4  1/2  I  6  1/8 
6.33       I        11.25     i        20.65 


4  1/2  Equation  (C) 
8  1/8  Equation  (D) 
36.6  Dwp  X  Lwp 


Pm 

250 

- 

Assumed 

Kwp 

1510 

1990 

2520 

2430 

2145 

Pm 

300 

Assumed 

Kwp 

1810 

2380 

3020 

2920 

2580 

Pm 

350 

Assumed 

Kwp 

2120 

2780 

3520 

3410 

3010 

Pm 

400 

Assumed 

Kwp 

2420 

3190 

4030 

3900 

3440 

Table  33F. — Wrist  or  Piston  Pin  Bearing  Pressures  for  Vertical  Stationary  Gas 

Engines. 


Tables  33 G  to  33L,  inclusive,  are  from  an  article  by  G.  W.  Lewis  and  A.  G. 
Kessler  published  in  the  American  Machinist.  They  give  the  maximum 
unit  pressures  and  rubbing   speeds  on  the   bearings  of  automobile   engines 


ALLOWABLE  PRESSURES 


75 


for  various  assumed  explosion  pressures  and  for  various  sizes  of  engine 
cylinders.  They  were  developed  from  an  investigation  of  some  30  Amer- 
ican automobile  engines  and  represent  average  practice. 

D    =  cylinder  diameter  in  inches, 

Z)c&  =  bearing  diameter  of  center  bearing  in  inches, 

Lcb  ==  bearing  length  of  center  bearing  in  inches, 

Xc6  =  maximum   unit  bearing  pressure  on  center  bearing  in  pounds 

per  square  inch, 
Pm  =  maximum  explosion  pressure  in  pounds  per  square  inch  of  piston 
face. 

Dcb=o.32D+o.3''    (5) 

Lcb  =2.^  Deb- 2.2"    (6) 


D 


4   1/2 


5  1/2 


Deb 
Lcb 
Acb 


1  9/16^         I  3/4 

2  3/16  2  3/4 
342              4-3 


I  7/8  2  1/16 

3   1/16        3  5/8 

5-75     I         7-5 


From  equation  (5). 
From  equation  (6). 
Deb  X  Lcb. 


Pm 

250 

Assimied. 

Kcb 

690 

615 

620 

575 

Pm 

300 

Assumed. 

Kcb 

830 

730 

750 

690 

Pm 

350 

Assumed. 

Kcb 

970 

855 

870 

810 

Pm 

400 

Assumed. 

Kcb 

HOC 

980 

1000 

920 

Table  33G. — Maximum  Unit  Bearing  Pressure  on  Center  Bearings  of  Automobile 

Engines. 


76 


BEARINGS  AND  THEIR  LUBRICATION 


D    =  cylinder  diameter  in  inches, 

Z)/6  =  bearing  diameter  of  front  bearing  in  inches, 

Ljb  =  bearing  length  of  front  bearing  in  inches, 

Z/6=  maximum  unit  bearing   pressure   on  front   bearing  in  pounds 

per  square  inch, 
Pm  =  Maximum  explosion  pressure  in  pounds  per  square  inch  of  piston 

face. 

DJb=o.s2D+o.y'     (7) 

Lfb  =  D/b+i  i/S''     (8). 


D 


Dfb 
Lfb 
Afb 


Pm 


4  i/2 


5  i/2 


1  9/i6  I 

2  I1/16 
4.2 


1  3/4 

2  7/8 
5.02 


I  7/8     '       2   1/16      From  equation  (7). 
3  3  3/16      From  equation  (8). 

5.63  6.6  Dfb  X  Lfb. 


250 


Assumed. 


Kfb 


Pm 


Kfb 


Pm 


560 


665 


595 


640 


670 


300 


Assumed. 


710 


775 


800 


350 


Assumed. 


Kfb 

780 

830 

900 

930 

Pm 

400 

Assumed. 

Kfb 

890 

945 

1030 

1075 

Table  33H. — Maximum  Unit  Bearing  Pressure  on  Front  Bearings  of  Automobile 

Engines. 


ALLOWABLE  PRESSURE 


77 


D     =  cylinder  diameter  in  inches, 

Drh  =  bearing  diameter  of  rear  bearing  in  inches, 

Lrb   =  bearing  length  of  rear  bearing  in  inches, 

Arb  =  projected  bearing  area  of  rear  bearing  in  square  inches, 

Krb  =  maximum  bearing   pressure   on  rear  bearing  in  pounds  per 

square  inch, 
Pm  =  maximum  explosion  pressure  in  pounds  per  square  inch  of  piston 

face. 
Drb=o.s2D  +0.3"   (9) 

Lrb^S-2>Drb-S-Z"     (10). 


4  1/2 


51/^ 


Drb 
Lrb 
Arb 


1-9/ 16 
3 
4-7 


1-3/4 


1-7/8         2-1/16 
4-5/8         5-5/8 
8.67         II. 6 


From  equation  (9) 
From  equation  (10) 
Drb  X  Lrb. 


Pm 

250 

Assumed. 

Krb 

56s 

495 

495 

465 

Pm 


Krb 


300 


Assumed. 


Pm 

350 

Assumed. 

Krb 

760 

660 

655 

605 

Pm 

400 

Assumed. 

Krb 

855 

735 

735 

675 

Table  33I. — Maximum  Unit  Bearing  Pressure  on  Rear  Bearings  of  Automobile 

Engines. 


78 


BEARINGS  AND  THEIR  LUBRICATION 


D       =  cylinder  diameter  in  inches, 

Dwp  =  bearing  diameter  of  wrist  pin  in  inches, 

Lwp   =  bearing  length  of  wrist  pin  in  inches, 

Awp  =  projected  bearing  area  of  wrist  pin  in  square  inches, 

Pm    =  maximum  unit  explosion  pressure  in  pounds  per  square  inch. 

of  piston  face. 

Dwp  =o.T,4D-o.sf    (i) 

Lwp   =2.2^  Dwp  (2) 

Kwp  =  maximum  unit  bearing  pressure  in  pounds  per  square  inch. 

Pm  J92  0.7854 

Kwp= -— {a) 

Awp 


Pm 


Kwp 


D 

4 

4  1/2 

5 

5  1/2 

Dwp 
Lwp 
Awp 

7/8 
2 

1-75 

I 

2.25 

2.25 

1  3/16 

2  5/8 
3.12 

I  3/8 
3  1/16 
4.2 

From  equation  (i). 
From  equation  (2). 
Dwp  X  Lwp. 

1800 


250 


1780 


1570 


1420 


Assumed. 


Equation  (a). 


Pm 


300 


Assumed. 


Kwp 


Pm 


2150 


2130  1890       j        1700  Equation  (a). 


350 


Assumed. 


Kwp 


2510 


2480  2200  1980       j   Equation  (a). 


Pm 

400 

Assumed. 

Kwp 

2870 

2840 

2510 

2270 

Equation  (a). 

Table  33  J. — ^Maximum  Unit  Wrist  Pin  Bearing  Pressures  in  Automobile  Engines. 


ALLOWABLE  PRESSURES 


79 


D      =  cylinder  diameter  in  inches, 

Dcp  =  bearing  diameter  of  crank  pin  in  inches, 

Lcp   =  bearing  length  of  crank  pin  in  inches, 

Acp  =  projected  bearing  area  of  crank  pin  in  square  inches, 

Kcp  =  maximum  unit  bearing  pressure  on  crank  pin  in  pounds  per 

square  inch, 
Pm  =  maximum  unit  explosion  pressure  in  pounds  per  square  inch 

of  piston  face. 

Dcp=  o.z2D  +  o.f         (3) 

Lcp  =  1.35  Dcp  (4) 

FmD' 0.JS54       _ 

^^^= A^p (') 


4-1/2 


S-1/2 


Dcp 
Lcp 
Acp 


1  9/16  I  3/4 

2  1/18         2  3/8 
3-32  4-17 


1  7/8 

2  1/2 
4.68 


2  1/16 

2  3/4 
5-68 


From  equation  (3) 
From  equation  (4) 
Dcp  X  Lcp. 


Pm 

250 

Assumed. 

Kcp 

945 

960 

1050 

1050 

From  equation  (b). 

Pm 

300 

Assumed. 

Kcp 

1 130 

1 1 50 

1260 

1260 

From  equation  (b). 

Pm 

350 

Assumed. 

Kcp 

1320 

1350 

1470 

1470 

From  equation  (b). 

Pm 

400 

Assumed. 

Kcp 

1510 

1540 

1675 

167s 

From  equation  (b). 

Table  33  K. — Maximum  Unit  Crank  Pin  Bearing  Pressures  for  Automobile  Engines. 


8o 


BEARINGS  AND  THEIR  LUBRICATION 


D         =  cylinder  diameter  in  inches, 

L  =  stroke  in  inches, 

■^mean  "^  Hican  total  prcssuie  on  piston  in  pounds  for  entire  cycle 

(assumed), 
RPM  =  revolutions  per  minute  at  looo  ft.  per  minute  piston  speed, 
Km      =  mean  unit  bearing  pressure  on  bearings  in  pounds  per  square 

inch, 
V         =  rubbing  speed  in  feet  per  second, 

W        =  work  of  friction  =  {Km  X  V/j.)  in  foot  pounds  per  second, 
L  =  1.1  D. 


D 


4-1/2 


5-1/2     }    Average 


4-3/8 


5-1/2 


R.P.M. 


[370 


1090 


Rubbing  speed  in  ft.  per  min.  on      560 
bearings  listed  below. 


550 


535 


540 


546 


V 


9-35 


915 


8.9 


252 


320 


395 


475 


Crank-pin  bearing. 


Center  bearing 


Front  bearing. 


Km 


76 


77 


84 


W 


710 


705 


Km 


55 


49.2 


W 


515 


450 


745 


49. 


445 


Km  44.7 


47-5 


52 


W 


418 


435 


462 


84 


755 


46 


415 


53-4 


480 


80 


729 


50 


456 


49-4 


449 


Rear  bearing. 


Km 


56 


45-5 


43-5 


38 


[  W 


523 


415 


388 


342 


45-7 


417 


Table  33L. — Average  Rubbing  Speed  and  Work  of  Friction  for  Automobile 
Engines. — Computed  for  a  piston  speed  of  1000  ft.  per  minute  and  a  mean  pressure  of 
20  lb.  per  square  inch  of  piston  face. 


ALLOWABLE  PRESSURES 


8i 


— 1 

r 

1 

_ 

\ 

\ 

\ 

\ 

'i 

1 

5 

\ 

J 

\ 

1 
C 

^ 

s§ 

fi 

p 

la 

V 

\ 

— 1 

^^ 

\ 

s 

\ 

^ 

•eajy  ps^oarojjj  qoui  gjBnbs  jad  spunoj  ui  oanssaa j  SuvvsdQ 


82  BEARINGS  AND  THEIR  LUBRICATION 

Fig.  II  gives  maximum  safe  bearing  pressures  for  rubbing  speeds  up  to 
5000  ft.  per  minute  and  for  perfect  film  lubrication.  It  shows  the  practice  of  the 
(jeneral  Electric  Company  for  the  bearings  of  electrical  machinery  and  is 
representative  of  the  best  of  American  practice. 

REPRESENTATIVE  GERMAN  PRACTICE 

Representative  German  practice  in  regard  to  unit  pressures  for  journal 
bearings  as  found  on  page  742  Des  Ingenieurs  Taschenbuch,  Hiitte,  Vol.  i,  is 
given  below  with  the  metric  units  transformed  into  corresponding  English 
units  in  round  numbers. 


Materials  in  contact 

Pressure  in  pounds  per 
square    inch    of    pro- 
jected area. 

Hardened  crucible  steel  on  hardened  crucible  steel 

2130 
1280 

Unhardened  crucible  steel  on  bronze 

850 

425 

355 
355 

Mild  steel  with  a  smooth  compact  surface  on  bronze 

Mild  steel  with  not  an  entirely  clear  surface  or  cast-iron  on  bronze. 
Mild  steel  with  not  an  entirely  clear  surface  on  cast  iron 

Mild  steel  on  lignum  vita?  with  water  lubrication 

In  explanation  of  the  above  values,  we  read:  ''If  the  total  pressure  even 
when  the  journal  is  at  rest  maintains  approximately  of  the  same  magnitude  and 
direction  (for  example,  in  heavily  loaded  shafts,  heavy  gears,  etc.)  the  unit 
pressure  should  be  taken  smaller  than  above. 

''For  journals  (or  bearings)  which  oscillate  only  the  unit  pressure  may  be 
considerably  higher. 

"For  the  bearings  of  rope  and  chain  sheaves,  etc.,  which  turn  intermittently 
and  on  which  the  wear  is  either  small  or  unimportant,  the  above  unit  values, 
850  to  355  lb.,  may  be  doubled  or  tripled. 

''For  crucible  steel  crank  and  cross-head  pins  running  on  bronze  in  ordinary 
steam  engines  the  unit  pressure  may  be  taken  from  710  to  995  and  1070  to  1130 
lb.,  respectively;  for  similar  purposes  on  locomotive  1420  and  21301b.,  respect- 
ively; on  high  speed  steam  engines  570  to  710  lb.,  respectively.  For  fly  wheel 
bearings  of  steam  engines  use  213  to  217  lb. 

"For  crank  pin  bearings  of  punches  and  shears  the  unit  pressure  may  be  as 
high  as  2845  It),  per  square  inch  of  projected  area." 

PERMISSIBLE  VELOCITY  OF  RUBBING 

There  are  no  well  defined  limits  for  rubbing  velocities.  In  many  classes  of 
machinery  the  speed  is  so  low  that  the  velocity  is  an  unimportant  factor.     With 


ALLOWABLE  TEMPERATURES  83 

the  exception  of  power  generating  machinery,  particularly  steam  turbines,  high 
speed  is  usually  accompanied  with  very  low  pressure.  For  instance,  in  machine 
tools  speeds  are  uniformly  low  with  the  exception  of  grinding  machinery,  and 
there  the  pressures  are  slight. 

In  Power,  February,  1893,  C.  J.  Field  gives  350  ft.  per  minute  as  the  max- 
imum surface  velocity  for  the  main  bearings  of  steam  engines  in  elect'-ical 
service  with  grease  lubrication.  The  corresponding  maximum  pressure  per 
square  inch  of  projected  area  of  journal  is  80  pounds. 

The  ordinary  range  of  velocities  for  electric  motors  and  generators,  as 
indicated  by  the  curve  in  Fig.  11,  is  from  400  to  1200  ft.  per  minute. 

Steam  turbine  practice  in  America  is  from  2400  to  6000  ft.  per  minute. 
However,  this  higher  figure  has  been  exceeded.  One  case  that  came  to  the 
author's  attention  was  of  a  horizontal  steam  turbine  with  a  bearing  rubbing 
velocity  of  8000  ft.  per  minute. 

PERMISSIBLE  RISE  IN  TEMPERATURE 

A  bearing  shows  distress  by  heating — by  becoming  a  ''hot  box."  Within 
limits  there  is  no  objection  to  the  rise  in  temperature;  in  fact,  it  must  take  place. 
It  is  sometimes  said  that  the  hotter  a  bearing  runs  the  hotter  it  can  run.  The 
reason  for  this  statement  lies  in  the  fact  that  the  higher  the  temperature  the  less 
the  viscosity  of  the  lubricant,  therefore  the  less  the  work  of  friction  and  the 
less  the  amount  of  energy  transformed  into  heat  at  a  given  speed. 

With  soft  metal  bearings  the  limiting  temperature  is  one  at  which  the 
lining  will  melt.  With  other  materials  the  limiting  temperature  is  one  at 
which  lubrication  becomes  insufficient  and  the  bearing  seizes. 

In  practice  it  is  necessary  to  design  bearings  to  run  at  a  much  lower  tem- 
perature than  will  cause  damage  because  of  the  requirements  of  the  average 
customer.  Such  a  maximum  temperature  is  from  140  to  160°  F.  It  is  prob- 
ably true  that  the  average  bearing  running  at  that  temperature  could  just  as 
well  run  at  a  temperature  of  200°  F.  Occasionally  manufacturers  take  advantage 
of  this  fact  in  designing  a  machine  to  be  located  in  some  out-of-the-way  place 
where  it  will  not  have  close  inspection,  or  will  not  be  approached  by  the  ordinary 
passer-by  who  might  place  his  hand  upon  a  bearing,  and  a  high  temperature  is 
allowed. 

There  is  no  valid  objection  to  running  a  bearing  hot,  provided  the  temper- 
ature reached  is  not  sufficient  to  do  harm  and  provided  there  is  no  progressive 
increase  in  heat  indicating  trouble.  This  is  opposed  to  the  ideas  of  some 
machinery  users  who  look  upon  any  bearing  that  warms  up  as  improperly 
designed.  There  is  no  reason  in  expecting  or  demanding  that  a  machinery 
bearing  should  run  stone  cold.  In  the  very  nature  of  the  device,  energy  must 
be  transformed  into  heat,  and  this  heat  will  cause  a  rise  in  the  temperature  of 


84 


BEARINGS  AND  THEIR  LUBRICATION 


the  bearing  until  a  condition  is  reached  where  the  heat  radiated  and  dissipated 
in  a  unit  of  time  balances  the  amount  liberated  in  the  same  time. 

In  machine-tool  practice  low  running  temperatures  are  the  rule.  This 
indicates  a  very  ample  factor  of  safety.  The  Bullard  Machine  Tool  Company 
has  fixed  the  maximum  running  temperature  under  load  for  bearings  of  the  ma- 
chines that  they  build  as  i  io°  F.  Every  machine  is  tested  in  this  particular  before 
shipment.  The  actual  running  temperatures  are  considerably  lower  as  shown 
by  the  following  tabulation,  giving  the  log  of  tests  of  six  machines. 


Machine 


24-inch   vertical 
turret  lathe. 


24-inch   vertical 
turret  lathe. 


24-inch   vertical 
turret  lathe 


36-inch   vertical 
turret  lathe. 


36-inch   vertical 
turret  lathe. 


36-inch   vertical 
turret  lathe. 


Running  Temperature  of  Bearings  in  Degrees  Fahrenheit 


Speed 
box 


Head- 
stock 


Table 
pinion 


f 

I-  70 

65 

J 

2-  80 

68 

3-85 

68 

I-  79 

61 

2-  95 

65 

3-  91 

67 

I-  74 

64 

2-  96 

68 

3-  95 

68 

r 

I--  75 

62 

1 

2-  81 

65 

3-80 

65 

r 

I-  82 

70 

2-100 

72 

3-  97 

72 

r 

I-  80 

64 

2-  85 

65 

^ 

3-  86 

66 

76 

79 
80 


Rail 
raising 
bracket 

73 

78 

80 

78 
82 
82 

85 
no 
112^ 


Power  travers- 
ing bracket 


Upper 


Lower 


74 
80 
80 


77 
80 

79 


86 
84 
82 

75 
80 
80 


68  70 

71  74 

72  74 


70 
76 
78 

76 

80 
78 


68  66 

70  70 

71  72 


85 
83 

74 
78 
79 


Feed  works 


Upper     Lower 


Driving 
pulley 
bracket 


66      '       65  70 

68  69  80 

68      1      69      I       78 


The  three  sets  of  readings  in  each  case  were  taken:  i,  at  the  end  of  the  first  hour's  run 
without  load;  2,  at  the  end  of  the  second  hour's  run  with  load;  3,  at  the  end  of  the  third 
hour's  run  without  load. 

^  After  correction  of  this  bearing  the  temperature  dropped  to  78°  at  the  end  of  a  2-hr.  run 


ALLOWABLE  PRESSURES  85 

PRODUCT  OF  UNIT  PRESSURE  AND  VELOCITY 

In  designing  bearings  neither  the  pressure  nor  the  circumferential  speed 
alone  can  serve  as  a  standard  for  calculating  dimensions.  The  product  of  the 
pressure  by  the  speed  and  the  coefficient  of  friction  must  be  taken  into  account, 
that  is,  the  work  of  friction. 

In  Section  II  it  is  shown  that  the  product  of  the  unit  pressure  and  the 
coefficient  of  friction  remains  almost  a  constant  for  a  wide  range  of  pressures. 
This  has  been  taken  advantage  of  by  many  designers  to  establish  a  limiting 
factor  of  design  commonly  referred  to  as  C,  which  represents  the  product  of 
the  pressure  in  pounds  per  square  inch  and  the  velocity  in  feet  per  minute. 

However,  the  values  of  this  factor  differ  so  widely  that  no  attempt  will  be 
made  here  to  tabulate  them  for  general  types  of  bearings.  To  give  an  idea  of 
the  range,  F.  W.  Taylor,  Trans.  A.S.M.  E.,  Vol.  XXVII,  gives  C  for  mill  work 
as  24,000  and  for  cast-iron  bearings  with  ordinary  lubrication  as  12,000.  The 
same  factor  used  in  turbo-generator  design  for  forced  lubrication  and  water 
cooling  is  360,000.  It  is  probable  that  these  may  be  taken  as  limiting  values, 
but  it  is  impossible  to  give  consistent  data  for  other  kinds  of  bearings  in  order 
to  complete  the  series. 

Tower  mentions  two  cases  in  which  his  experimental  bearing  seized,  the 
conditions  being  a  bronze  half  box,  a  steel  journal  and  bath  lubrication.  In 
one  instance  the  bearing  seized  at  a  pressure  of  573  lb.  per  square  inch  and  a 
speed  of  419  ft.  per  minute.  This  is  equivalent  to  a  factor  of  240,000.  In 
the  other  case  the  pressure  was  625  lb.  per  square  and  the  rubbing  velocity 
366  ft.  per  minute.     The  corresponding  factor  is  228,000. 

One  instance  that  has  come  to  the  author's  attention  is  of  a  large  turbo- 
generator running  with  a  bearing  pressure  of  41  Ib-per  square  inch  and  rubbing 
velocity  of  8,000  ft.  per  minute.     The  value  of  C  in  this  case  is  328,000. 

Some  of  the  greatest  values  for  this  factor  that  have  come  to  the  author's 
attention  are  from  some  experiments  made  by  the  General  Electric  Company. 
In  one  case  a  bearing  was  run  at  a  pressure  of  285  lb.  per  square  inch  and  a 
rubbing  velocity  of  2,700  ft.  per  minute — C=  770,000.  In  the  second  case  the 
pressure  was  200  lb.  and  the  velocity  5,000  ft. — C=  1,000,000. 

The  Westinghouse  Machine  Company  has  carried  this  experimentally  to 
1,720,000;  400  lb.  pressure,  4300  ft.  per  minute,  and  for  long  runs. 

In  spite  of  the  fact  that  it  is  impossible  to  give  values  for  this  factor  that 
seem  to  be  consistent  with  themselves  and  in  sufficient  number  to  cover  the 
ordinary  range  of  commercial  bearings,  pressures,  and  speeds,  the  designer 
should  not  think  it  is  of  no  value.  On  the  contrary,  it  is  of  considerable  value 
as  a  rough  check  upon  design.  It  is  wise  for  any  one  having  to  do  with  bearing 
design  to  tabulate  this  factor  carefully  for  all  important  bearings  coming  within 
his  experience,  to  arrive  at  limiting  values  beyond  which  he  will  know  that  the 
running  of  the  bearing  is  liable  to  be  unsatisfactory  if  not  impossible. 


SECTION  V 

DESIGN  OF  JOURNAL  BEARINGS 

The  important  points  to  consider  in  proportioning  and  designing  journal 
bearings  may  be  grouped  under  six  heads:  - 

a.  Proportions  and  dimensions  of  diameter  and  length. 

b.  Selection  of  the  bearing  material. 

c.  Provisions  for  anchoring  the  bearing  metal  if  a  soft  lining,  and  provisions 
or  adjustment  to  preserve  fit  and  alinement. 

d.  Clearance  between  journal  and  bearing. 

e.  Provisions  for  lubricating. 

f.  Provisions  for  radiating  and  dissipating  heat. 

RATIO  OF  BEARING  LENGTH  TO  DIAMETER 

In  proportioning  bearings,  the  designer  must  harmonize  conflicting  con- 
ditions. On  the  one  hand  the  shaft  must  be  sufficiently  strong  and  rigid  to 
sustain  its  load  without  undue  deflection  (this  calls  for  large  diameters) ;  again 
the  bearing  pressures  must  not  exceed  a  certain  amount  in  order  that  an  oil 
film  may  be  established  and  maintained  (this  calls  for  large  diameters  and 
long  bearings).  On  the  other  hand  the  circumferential  speed  must  be  kept  as 
low  as  possible,  thus  keeping  down  the  work  of  friction  and  reducing  the 
amount  of  heat  liberated  and  the  risk  of  seizing  (this  calls  for  small  diameters). 
Again  the  bearing  length  should  be  short,  thus  rendering  the  surfaces  more 
effective  and  minimizing  the  danger  of  a  local  concentration  of  pressure  from  the 
deflection  of  the  journal  (this  calls  for  short  bearings). 

In  actual  designing  the  total  load  to  be  carried  by  the  bearing  divided  by 
the  allowable  unit  pressure,  gives  at  once  the  projected  area  of  the  bearing. 
The  smallest  permissible  diameter  of  journal  can  then  be  found  by  deter- 
mining the  shaft  deflection  under  running  conditions.  This  investigation  can 
best  be  made  by  graphic  statics.  A  valuable  discussion  of  this  subject  will  be 
found  in  Elements  of  Graphic  Statics,  by  Cathcart  and  Chaffee,  at  page  269. 
The  diameter  thus  determined  can  be  checked  by  finding  out  if  its  circumfer- 
ential rubbing  velocity  is  greater  than  allowable. 

If  the  first  proportions  selected  are  found  to  be  beyond  the  limits  of  good 
practice,  another  set  can  be  taken  and  the  investigation  repeated  until  the 

86 


DESIGN  OF  JOURNAL  BEARINGS  87 

proportions  are  deemed  satisfactory.  Section  IV  gives  allowable  bearing 
pressures  and  velocities. 

It  is  seen  from  the  preceding  principles  that  the  diameter  and  length  of  a 
bearing  should  both  be  kept  as  small  as  possible,  consistent  with  the  necessary 
strength  and  stiffness;  the  same  conclusion  is  reached  from  the  viewpoint  of 
economy  of  material. 

Practice  has  tended  to  standardize  the  ratio  of  bearing  length  to  diameter 
and  the  following  tabulation  gives  representative  proportions.  Here  I  is  the 
bearing  length  and  d  the  diameter. 


Type  of  bearing 


Values  of  -^ 
a 


I 

Marine  engine  main  bearings '  i  to  i .  5 

Marine  engine  crank  pin  bearings j  i  to  i .  5 

Stationary  engine  main  bearings i .  5  to  2 . 5 

Stationary  engine  crank  pin  bearings i 

Stationary  engine  cross-head  pin  bearings i  to  i .  5 

Ordinary  heavy  shafting  with  fixed  bearings 2  to  3 

Ordinary  shafting  with  self-adjusting  bearings 3  to  4 

Generator  and  motor  bearings i  2  to  3* 

Machine  tool  bearings |  2  to  4 


PROPORTIONS    OF  AUTOMOBILE   CRANK   SHAFT   BEARINGS 

A.  G.  Kessler  and  G.  W.  Lewis,  in  the  American  Machinist  of  September  29, 
1 9 10,  give  the  results  of  an  investigation  of  the  proportions  of  some  30  repre- 
sentative American  automobile  engines. 

The  diameter  of  the  front  crank  shaft  bearing  was  found  to  vary  from  i  1/2 
to  2  1/4  in.,  and  is  found  from  the  equation: 

where      -S^  =  diameter  of  the  front  bearing  in  inches, 
and  D  =  diameter  of  the  engine  cylinder  in  inches. 

The  length  of  the  front  bearing  is  expressed  by  the  equations: 
Zi=5i+  I  1/8  in.f  and 
L^=  1.67  ^1  +  0.95  in. 
The  reason  for  two  equations  lies  in  the  arrangement  of  the  curves  drawn 
through  the  points  plotted  from  the  dimensions  of  the  engines  investigated.     It 
shows  that  there  are  two  types  of  shaft  in  use. 

The  diameter  B2  of  the  rear  crank  shaft  bearing  is  found  from  the  equation: 

B2=o.S3D-s/Sin. 
This  is  identical  with  the  equation  for  the  diameter  of  the  front  bearing,  as 


88  BEARINGS  AND  THEIR  LUBRICATION 

it  is  customary  to  make  all  bearing  diameters  the  same  on  automobile  engine 
crank  shafts. 

The  length  of  the  rear  bearing  is  found  from  the  equation: 

^2=5-33  A- 5-66  in. 

2^2=  Length  of  rear  bearing  in  inches 
and  D=  diameter  of  engine  cylinder  in  inches,  as  above. 

SELECTION  OF  THE  BEARING  METAL 

The  bearing  metal  must  be  selected  with  reference  to  the  metal  in  the 
journal  with  which  it  is  to  run,  to  the  load  which  it  is  to  carry — it  must  be 
strong  enough  to  resist  deformation — and  in  many  classes  of  machinery  it 
must  be  selected  for  its  resistance  to  wear.     See  Section  III, 

Bronze  is  ordinarily  used  in  the  form  of  a  sleeve  or  bushing  either  in  a  single 
piece  or  split  in  half.  If  a  soft  metal  lining  (babbitt)  is  used,  it  may  either  be 
poured  directly  into  the  housing  of  the  bearing  or  inserted  in  a  removable  shell. 
Such  shells  are  commonly  made  of  cast  iron,  although  brass  is  used  to  a  limited 
extent  and  cast  steel  still   more  rarely. 

Standard  American  railway  practice  consists  in  alloying  the  lining  and  its 
bronze  shell  at  their  joining  surfaces.  The  bronze  surface  is  carefully  cleaned, 
coated  with  solder,  and  the  lining  poured  against  it.  If  the  work  is  well  done 
there  is  no  danger  of  the  lining  loosening.     There  is  no  mechanical  anchorage. 

Sometimes  different  metals  are  used  for  the  lower  half  of  a  bearing  and  its 
cap.  If  the  cap  takes  no  load  it  can  be  lined  with  a  cheaper,  softer  metal  than 
the  load-carrying  box. 

ANCHORING  OF  THE  METAL  LINING 

Babbitt  linings  must  be  securely  anchored  to  their  shells  for  if  they  loosen 
they  are  apt  to  crack  and  be  destroyed.  If  this  does  not  happen,  the  slight 
motion  between  the  lining  and  its  shell  may  act  as  pump  and  take  the  oil  out  of 
the  bearing.  The  devices  for  anchoring  are  simple.  In  a  common  form  the 
shell  has  a  series  of  ridges  with  dovetailed  edges  around  which  the  metal  is  cast. 

Fig.  12  shows  an  anchoring  arrangement  using  anchors  having  a  circular 
cross-section  from  the  practice  of  the  Westinghouse  Electric  and  Manufacturing 
Company.  The  patterns  are  made  to  leave  a  plain  core  in  the  green  sand,  to 
which  are  added  baked  anchor  cores  secured  with  brads  as  shown  .  There  are 
two  sizes,  1/2  and  3/4  in. 

In  bearings  that  for  any  reason  have  to  be  bored  before  babbitting,  the 
anchor  holes  are  cast  extra  deep,  so  that  after  boring  the  holes  will  be  standard 
depth. 


DESIGN  OF  JOURNAL  BEARINGS  89 

The  illustration  gives  the  spacing.  Along  the  straight  lips  of  each  half,  the 
anchors  should  be  fairly  close  together,  and  as  near  the  edge  as  they  can  be 
cast,  so  that  when  the  casting  is  finished  the  anchor  holes  will  break  through. 


Baked  Anchor  Core 


Center  Mark 


Standard  Anchors. 


Far  all  Split  Bearings 
less  thau  9"in  Dia. 


?<' Anchor 


-^  j;^//////////////, 


For  all  Split  Bearing 
g'or  more  in  Dia. 


Extra  Deep  Anchors. 


i 

^ 

i^/Ayy^^^'MX 

n^^ 

'^.'-v-? 

]-,,,._ 

1 

-''^;»_", 

"-' 4' f '-'-'' 

-/'./'-'O 

b 

^^^ 

^^^ 

^^^ 

These  Holes  (through  which  the  Babbitt 
is  poured)  are  Cast  in  the  Bottom  Half 
of  the  Bearing  directly  opposite  the  Oil 
Holes  and  Slots  in  the  Top  Half. 


'vi  under  5  Dia. 
ijafor  5"Dia.  and 


Pitch  P  for  Yi.  Anchor  about  \M. 
Pitch  P  for  ?4 "Anchor  about  \M'' 
Over  the  remainder  of  the  Bear- 
ing the  Holes  should  be  more 
sparsely  distributed.  „ 

Pitch  D  for  i^'Anchor  about  2M. 
Pitch  D  for  ?i"Anchor  about  3." 


FIG.    12.       SIZE   AND    PITCH   OF   BABBITT   ANCHORS. 


ADJUSTING    DEVICES    TO    PRESERVE   ALINEMENT 


Various  devices  are  used  to  preserve  alinement,  or  to  make  bearings  self 
alining  within  small  limits.  In  some  cases,  as  for  instance  the  spindle  bearings 
of  machine  tools,  the  accuracy  of  the  seat  controls  the  alinement.  In  small  elec- 
tric motors  and  a  large  number  of  shoe  machines  bushings  are  forced  into  place 
in  the  frames  or  housings  and  are  then  carefully  line  reamed.  In  case  a  bushing 
has  to  be  replaced  for  any  cause  whatever  the  process  of  line  reaming  must  be 
repeated. 

In  steam  engine  work  the  bearing  brasses  are  usually  adjustable  by  means 
of  wedges  or  bolts.     In  electric  motors,  generators,  horizontal  turbines,  and 


90 


BEARINGS  AND  THEIR  LUBRICATION 


some  gas  engines,  the  bearings  are  made  with  a  spherical  seat,  as  shown  by 

Fig.  13. 

Bearings  for  line  shafting  are  usually  held  on  a  self-alining  ball  seat  pro- 
vided with  adjusting  screws  in  a  vertical  direction  and  with  elongated  slots  in 
the  feet  of  the  hanger  for  sidewise  adjustment.  (See  Fig.  13.)  The  self  alining 
feature  lies  in  the  ball  seat  of  the  box. 

Machine  countershaft  bushings  and  occasionally  more  important  bearing 
linings  are  frequently  placed  in  position  on  the  shaft  inside  of  the  housings  and 


'^ 

r 

Ball  or  Spherical  Surface 
Electric  Motor  Bearing. 


Plain  Ball  and  Socket  Shafting  Box  (Sellers.) 
FIG.    13.      TYPICAL   SPHERICAL  ADJUSTING   SURFACES   FOR  BEARINGS. 

the  space  between  bushing  and  housing  run  with  babbitt  or  an  alloy  that  ex- 
pands slightly  in  cooling.  This  is  a  simple,  cheap,  and  satisfactory  manner 
to  aline  and  locate  such  bushings  having  all  of  the  advantages  in  manipulation 
possessed  by  the  babitted  box. 

On  some  textile  machinery  provided  with  rigid  bearings,  a  slight  amount 
of  self-adjustment  is  obtained  by  tapering  the  journal  from  full  diameter  at 
about  its  centre  to  a  reduction  of  0.003  or  0.004  in.  at  the  ends.  This  double 
taper  allows  for  a  slight  amount  of  swiveling. 


CLEARANCE  BETWEEN  BEARING  AND  JOURNAL 

The  space  or  clearance  between  the  rubbing  surfaces  is  of  importance. 
If  too  small  the  coefficient  of  friction  will  be  very  high  with  a  tendency  to  heat. 
If  too  large  it  will  be  difficult  to  maintain  a  suitable  film  of  oil.  The  shaft 
should  be  freely  supported  and  freely  guided  in  its  bearings,  without  producing 
any  local  strain  in  the  material  and  no  local  heating  through  shocks.  As  a 
rule  the  tendency  is  to  make  the  fit  too  close.     If  for  any  reason  a  journal 


DESIGN  OF  JOURNAL  BEARINGS 


91 


must  be  tightly  fitted  in  its  bearings,  which  means  that  the  entire  circumference 
of  the  surface  is  loaded  approximately  equally,  oil  should  be  supplied  by  forced 
lubrication.     This  condition  gives  a  high  amount  of  friction. 

Lasche  found  that  the  friction  force  decreased  rapidly  with  an  increasing 
clearance  becoming  practically  constant  when  the  clearance  was  0.06  in.  His 
experiment  was  with  a  bearing  10  in.  in  diameter  and  4  in.  long.  This  is  an 
extreme  case. 

Good  operating  practice  for  main  engine  bearings  is  a  slackness  of  0.005 
in.  for  diameters  from  6  to  12  in.  and  for  cross-head  pins  and  crank  pins  up  to 
6  in.  in  diameter  0.004  in. 

L.  D.  Burlingame,  Trans.  A.S.  M.  E.,  1910,  gives  these  limits  used  by  Brown 
&  Sharpe  Mfg.  Co.  for  ground  cylindrical  fits  of  small  size.  The  tolerances 
are  in  decimals  of  an  inch. 


Running  Fits — Ordinary  Speed 

To  1/2   -in.  diameter,  inc 0.00025  to  0.00075  Small 

To  I        -in.  diameter,  inc 0.00075  to  0.0015  Small 

To  2        -in.  diameter,  inc 0.0015    to  0.0025  Small 

To  3  1/2-in.  diameter,  inc 0.0025    to  0.0035  Small 

To  6        -in.  diameter,  inc 0.0035    to  0.005  Small 

Running  Fits — High  Speed,  Heavy  Pressure  and  Rocker  Shafts 

To     1/2-in.  diameter,  inc o  .0005  to  o  .001  Small 

To  I        -in.  diameter,  inc o.ooi    to  0.002  Small 

To  2        -in.  diameter,  inc 0.002     to  0.003  Small 

To  3  1/2  in.  diameter,  inc 0.003     t^  0.0045  Small 

To  6        -in.  diameter,  inc 0.0045  to  0.0065  Small 

Sliding  Fits 

To     1/2-in.  diameter,  inc o  .00025  to  o  .0005  Small 

To  I        -in.  diameter,  inc o .  0005  to  o .  001  Small 

To  2        -in.  diameter,  inc o.ooi     to  0.002  Small 

To  3  1/2-in.  diameter,  inc 0.002     to  0.0035  Small 

To  6        -in.  diameter,  inc o  .003     to  o  .005  Small 

Grinding  Limits  for  Holes 

To  I        -in.  diameter,  inc Standard  to  0.00075  Large 

To  2        -in,  diameter,  inc Standard  to  o  .001  Large 

To  3  1/2-in.  diameter,  inc Standard  to  0.0015  Large 

To  6        -in.  diameter,  inc Standard  to  o  .002  Large 

To  12      in.    diameter,  inc Standard  to  0.0025  Large 

Table  34  gives  the  standard  practice  of  the  General  Electric  Company. 


92 


BEARINGS  AND  THEIR  LUBRICATION 


A 

Journal 

Bearings 

Axle 
for  ry 

linings 

-.M 

Horizontal 

Vertical 

Step 

.  motors 

Allowable 

1 

•5   a 

^  i 

''5 

Max. 

diam. 

in. 

variation 
below 
max. 

diameter 

Min. 
bore 
in. 

Allowable 
variation 

above 
min.  bore 

Min. 
bore 
in. 

Allowable  j 
variation 

above 
min.  bore 

Min. 
bore 
in. 

Allowable 
variation 

above 
min.  bore 

Min. 
bore 
m. 

Allowable 
variation 

above 
min.  bore 

3/8 

.375 

.0005 

.377 

.001 

.376 

.001 

.3755 

.0005 

.380 

.004 

1/2 

.500 

0005 

.502 

.001 

.501 

.001 

.5005 

.0005 

.505 

.004 

5/8        .625 

.0005 

.627 

.001 

.626 

.001 

-6255 

.0005 

.63c 

.004 

3/4 

.750 

.0005 

.752 

.001 

.751 

.001 

.7505 

.0005 

.755 

.004 

7/8 

.875 

.0005 

.877 

.001 

.876 

.001 

.8755 

.0005 

.880 

.004 

I 

1 .000 

.0005 

1.002 

.001 

1 .001 

.001 

I .0005 

.0005 

1.005 

.004 

I  1/8 

1. 125 

.0005 

1. 128 

.001 

1. 127 

.001 

1 .  126 

.0005 

1 .  13c 

.004 

I   1/4 

1.250 

.0005  • 

1.253 

.001 

1.252 

.001 

I. 251 

.0005 

I-2S5 

.004 

I  1/2 

1 .500 

.0005 

1.503 

.001 

1 .502 

.001 

1.501 

.0005 

I-S05 

.004 

I  3/4 

1.750 

.0005 

1.753 

.001 

1.752 

.001 

I. 751 

.C005 

I. 755 

.004 

2 

2.000 

.0005 

2.003 

.001 

2 .002 

.001 

2.001 

.0005 

2.005 

.004 

2   1/4    2.250 

.0005 

2.253 

.001 

2.252 

.001 

2.251 

.0005 

2-255 

.004 

2   1/2    2.500 

.0005 

2.503 

.001 

2.503 

.001 

2.501 

.0005 

}  2.505 

.004 

2  3/4    2.750 

.0005 

2.754 

.002 

2.753 

.002 

2.7515 

.0005 

2.75s 

.004 

3              3.000 

.0005 

3.004 

.002 

3.003 

.002 

3.0015 

.0005 

3 -00s 

.004 

3   1/2    3.500 

.001 

3-504 

.002 

3.504 

002 

3.5015 

.0005 

3.507 

.004 

4             4.000 

.001 

4.005 

.002 

4.004 

.002 

4.002 

.001 

4-007 

!       .004 

4   1/2:   4.500 

.001 

4-505 

.002 

4.504 

.002 

4.502 

.001 

4-509 

.004 

5         1   5.000 

.001 

5.006 

.002 

5.005 

.002 

5. 0025 

.001 

:  5-009 

.004 

5   1/2    5.S00 

.001 

5.507 

.002 

5 -505 

.002 

5.503 

.001 

5-511 

.004 

6 

6  .000 

.001 

6.009 

.002 

6.005 

.002 

6.003 

.001 

j   6. on 

.004 

7 

7.000 

.001 

7.011 

.002 

7.006 

.002 

7.0035 

.CCI 

i    7.012 

.004 

8 

8.000 

.001 

8.012 

.003 

8.006 

.003 

8.004 

.C02 

8.013 

.004 

9 

9.013 

.004 

y.       / 

.004 

9-0045 
10.005 
11.0055 
12.006 

12 

12.000 

.0015 

12 .016 

.005 

12,008 

.005 

.002 

1 

13  :r3.ooo 

14  14.000 

15  It;.  000 

.0015 

13.016 
14.016 

.  005 

13.009 
14.009 
15 .010 

.005 

13.0065 

14.007 

iS-0075 

16.008 

002 

.0015 

.005 

.  005 

002 

.0015 

IS .016 

.005 

.005 

• 
.  002 

1 

16 

16.000 

.0015 

16.016 

.005 

16.010 

.005 

.002 

17 

1 7 . 000 

.0015 

17.018 

.005 

17. on 

.005 

17.008 

18 

18.000 

.0015 

18.018 

.005 

18. on 

.005 

18.008 

.002 

19 
20 

19.000 
20.000 

.0015 
.0015 

19.018 
20.018 

.005 
.005 

19.012 
20.012 

.005 
.005 

19.008 
20.008 

002 

.002 

21 

21 .000 

.002 

21.018 

.005 

21.013 

.005 

21.008 

.002 

22 

22.000 

.002 

22.02c 

.008 

22.013 

.005 

22.008 

.002 

i 

.  -    oon 

23 .02c 
24 .020 
25.020 
26.020 
27.022 
28.022 
29.022 

23-013 
24-013 

.005 
.005 

23 .008 
24.008 

■^         ,--• 

.008 

25 
26 

25.000 
26.000 
27.000 
28.000 
29.000 

.003 
.003 
.003 
.003 
.003 

.oc8 

.008 
.008 

j 

27 
28 



j        ! 

.008 
.008 

29 

30 
31 

30.000 
31 .000 

.003 
.003 

30.022 
31 .022 

008 

.008 



1 

j 

32 
33 
34 
35 
36 

32.000 
33.000 
34.000 
35- 000 
36.000 

.003 
.003 
.003 
.003 
.003 

32.024 
33.024 
34-024 
35.024 
36.024 

.010 

; 

1 

... 

i 

i::::::::: 

i 

1 

.010 

!■•;■:■ 

1 

1 

Table  34. — Allowances  and  Limits  for  Journals  and  Bearings. 


DESIGN  OF  JOURNAL  BEARINGS 


93 


PROVISIONS  FOR  LUBRICATING 

American  lubricating  oils  and  greases  are  by  no  means  standardized  as  to 
quality  and  constituents.  It  is,  therefore,  necessary  for  the  designer  to  con- 
sider the  possibility  of  a  very  poor  lubricant  being  used,  and  prepare  for  that 
condition.  American  machinery  builders  do  not  specify  what  oil  shall  be  used 
except  by  the  expression  ''a  good  oil  for  the  purpose." 

Therefore,  at  the  outset,  the  designer  must  prepare  for  the  contingency  of  a 
poor  oil  or  grease. 

As  to  the  method  of  applying  the  lubricant  there  is  a  wide  choice  and  it  is  one 
of  the  most  important  features  of  design.  At  the  same  time  it  is  one  that  is 
ordinarily  neglected;  even  a  superficial  reading  of  the  pages  of  technical  peri- 
odicals will  show  many  protests  on  the  part  of  users  of  machinery  against 
inadequate  oiling  devices;  even  the  elementary  principles  of  feeding  oil  to  a 
bearing  are  ofttimes  ignored.  On  page  1 8  it  is  pointed  out  that  a  properly  fitted 
and  designed  bearing  will  suck  oil  against  a  considerable  head,  provided  it  is 
introduced  at  a  proper  point.  It  has  also  been  pointed  out  that  the  oil  film  in  a 
well  lubricated  bearing  is  under  considerable  pressure.  Obviously  oil  by  a 
gravity  feed  cannot  enter  a  bearing  at  this  latter  point.  As  a  general  principle 
oil  fed  by  gravity  must  be  introduced  into  a  bearing  at  about  the  point  of  least 
pressure. 

The  common  methods  may  classified  as  follows: 

From  a  squirt  can  through  an  open  oil  hole. 
From  a  squirt  can  upon  a  felt  washer  covering 

an  oil  hole. 
From  a  squirt  can  through  an  oil  hole  cover 

to  the  oil  hole  beneath. 

Oil "I   From  an  oil  cup  through  a  wick  (syphon  feed). 

From  a  sight-feed  oil  cup. 

From  a  pad  pressed  against  the  journal. 

From  an  oil  well  by  moving  rings  or  chains  or 

by  vanes  fastened  to  the  journal. 
From  an  oil  well  by  centrifugal  means. 


Ordinary  or  imperfect  lubrica- 
tion. 


{From  a  compression  grease  cup. 
From  a  stick  of  grease  pressed  against  the 
journal. 


Perfect  lubrication Oil. 


From  a  pump  or  gravity  system  under  pres- 
sure (15  pounds  up)  (the  oil  may  or  may 
not  be  artificially  cooled). 

From  a  bath  in  which  the  journal  dips. 

From  a  reservoir  in  which  the  bearing  is  sub- 
merged or  from  which  it  receives  a  flood 
under  very  little  pressure  (2  to  5  pounds). 


94 


BEARINGS  AND  THEIR  LUBRICATION 


The    relative    values    of    several  methods  of  lubrication  were  found    by 
Goodman,  Proc.  Inst.  C.  E.,  Vol.  LXXXIX,  page  447,  to  be: 


Bath 

Saturated  pad 

Ordinary  pad 

Syphon 

I               

1.32 

2.21 

4. 2 

The  numerical  quantities  indicate  the  frictional  resistances  referred  to  the 
best — bath  lubrication — as  i. 

For  all  gravity  feeds,  the  point  of  application  should  be  at  about  the  point  of 
minimum  pressure  in  the  oil  film. 

With  forced  feed,  practice  indicates  that  the  point  of  application  should  be  at 
an  angular  distance  of  45  degrees  from  the  vertical  in  a  direction  counter  to  the 
direction  of  rotation.     This  is  indicated  by  the  diagram  Fig.  14. 


Point  to  Introduce 

Oil  under  Pressure     ^  / 

FIG.    14.      LOCATION   TO    INTRODUCE    OIL   WITH    FORCED    LUBRICATION. 


DISCUSSION  OF  THE  METHODS 

The  common  open  oil  hole  has  a  serious  objection;  it  is  an  opening  to  admit 
dirt  and  grit  into  the  bearing  as  well  as  oil.  With  the  felt  plug  it  is  consider- 
ably improved  as  dirt  is  excluded  and  the  felt  acts  as  a  small  reservoir  from  which 
oil  can  drip.  (See  Fig.  15.)  Many  self-closing  oil  hole  covers  are  on  the  market 
and  serve  their  purpose  until  they  are  broken  off  or  lost.  On  some  kinds  of 
machinery  they  are  used  as  ornaments. 

The  oil  pad  in  contact  with  the  journal  is  a  universal  method  for  lubricat- 
ing railroad  car  bearings  in  Europe,  and  is  extensively  used  for  line  shafting. 
It  is  open  to  the  objection  that  in  time  it  will  become  glazed  on  the  surface  in 
contact  with  the  moving  journal  and  lose  its  capillarity  and  ability  to  transfer  oil 
from  the  well  below  to  the  journal  above.  A  remedy  is  to  take  it  out  and  soak 
it  in  kerosene  or  gasoline  to  remove  the  accumulation  of  dirt. 


DESIGN  OF  JOURNAL  BEARINGS 


95 


The  pad  itself  is  often  made  of  a  fabric  having  a  face  of  wool  velvet  and  a 
back  of  cotton.  This  is  fastened  to  a  block  of  wood,  and  pressed  against  the 
journal  by  springs  which  lead  down  to  the  oil.  In  some  designs,  narrow  ledges 
of  the  wood  back  fit  the  journal  to  reduce  the  wear  on  the  pad.  The  arrange- 
ments are  satisfactory  but  are  open  to  two  objections  in  America,  the  expense 
and  difficulty  of  adapting  them  to  standard  equipment. 

The  standard  method  of  lubricating  American  car  axle  journals  is  by  wool 
waste  packing.  The  wool  is  elastic  even  when  oil-soaked.  It  has  a  disadvan- 
tage of  less  capillary  power  than  cotton  and  of  short  length  fibers.  Cotton, 
however,  loses  its  elasticity  when  wet. 

Oiling  Hole  %.  Vz  or 
%  Inch  in  Diameter. 

/ TTTI 


FIG.    15.       OPEN   OILING  HOLE    WITH    FELT   WASHER. 


Recent  experiments  have  been  made  with  a  woven  fabric  supported  close  to 
the  edges  of  the  brass  and  hanging  down  in  a  catenary  form  into  the  oil.  The 
pressure  against  the  journal  is  very  slight,  and  long  wicks  bring  up  the  oil.  The 
results  are  promising. 

Ring  or  chain  lubrication  is  the  preferred  form  for  better  classes  of  machin- 
ery where  the  heat  from  the  bearing  can  be  dissipated  without  artificial  means 
as,  for  instance,  in  electric  generators  and  motors.  Fig.  16  is  representative  of 
the  arrangement.  The  rings  hang  loosely  on  the  journal,  rotate  with  it,  pass 
around  the  outside  of  the  box,  and  dip  into  a  reservoir  of  oil  beneath;  thus  they 
move  in  the  same  direction  as  the  journal  and  carry  oil  to  its  upper  surface. 
A  certain  amount  finds  its  way  to  the  rubbing  surfaces  and  lubricates  the  bear- 
ing. The  rest  is  merely  transferred  from  one  side  of  the  reservoir  to  the 
other. 

Lasche  states  that  there  is  a  limit  of  speed  beyond  which  rings  cannot  be 
used  successfully,  commonly  said  to  be  the  speed  at  which  they  begin  to  ''  dance  " 
on  the  shaft  instead  of  rotate.  In  his  experiments  he  found  that,  considering  the 
amount  of  oil  delivered  to  the  bearings,  this  maximum  speed  was  about  2,000 
r.  p.  m.,  and  at  about  2,500  revolutions  the  amount  of  oil  delivered  decreased 
very  rapidly  with  a  further  increase  of  speed. 


96 


BEARINGS  AND  THEIR  LUBRICATION 


American  practice,  however,  uses  rings  at  journal  speeds  as  high  as  3,600 
r.  p.  m.  The  "dancing"  is  stopped  by  iacreasing  the  weight  of  the  ring  and 
depth  of  submersion.  After  a  journal  speed  of  from  600  to  800  r.  p.  m.  has 
been  reached,  the  rings  do  not  increase  in  speed  for  an  increase  of  journal 
speed.  Good  proportions  are  a  ring  diameter  equal  to  two  journal  diameters 
and  a  submersion  in  the  oil  equal  to  one-half  the  journal  diameter.  Allowing 
for  a  suitable  clearance  around  the  rings  and  the  size  and  capacity  of  the  oil 
well  are  determined. 

The  curves  of  Fig.  1 7  are  from  Lasche's  experiments  and  may  be  used  as  a 
guide  in  estimating  the  amount  of  oil  delivered  by  rings  to  the  rubbing  surfaces 
of  the  bearings. 


Housing 


Conical  Collar  to 
return  Oil  to 
Reservoir 


FIG.    16.       RING   OILING   BEARING. 


These  curves  are  especially  remarkable  in  that  they  show  how  widely  differ-^ 
ent  quantities  of  oil  are  delivered  for  similar  bearing  conditions.  When  the 
journal  revolved  in  the  direction  of  the  hands  of  a  watch,  the  quantity  of  oil 
flowing  was  only  1/2  to  2/3  of  that  which  flowed  when  the  journal  revolved  in 
the  opposite  direction.  Further,  in  both  directions  of  rotation  the  quantity  of 
oil  which  flowed  from  the  farther  side  was  only  about  1/3  of  that  which  escaped 
from  the  side  nearer  the  driving  belt.  When  the  journal  revolved  in  the  more 
favorable  direction  and  the  oil  entered  at  the  unloaded  side,  0.19  pint  per  minute 
flowed  through  the  bearing.  When  it  revolved  in  the  other  direction  the  quantity 
was  only  0.12  pint  per  minute. 

That  the  relationship  of  these  quantities  must  not  be  taken  as  indicating  a 


DESIGN  OF  JOURNAL  BEARINGS 


97 


law  is  shown  by  the  second  bearing  tested  in  which  the  quantities  of  oil,  de- 
pending upon  the  direction  of  rotation,  happened  to  be  in  a  reverse  order. 
To  work  well  the  rings  should  be  smooth  and  free  from  sharp  edges  and 


0  200  500      1000      1500     2000     2500 
Eevolutions  per  Minute 


Direction  of_ 
Rotation  IIT 


Direction  of 
Rotation  L.~ 


0.20 
»  0.189  «  0.18 
a  0.168  i^  O.IC 
O-cj  0.147  2«J0.14 
c  ^  0.12G  .S  1 0.12 
'^^  0.105  >.=  0.10 
g^  0.084  g  1, 0.08 
3  o.  0.063  =  P.O.OC 
"  0.042  -  0.04 
5  0.021  b  0.02 
0  0 


Total  iQuantityofI  Oil  Flowing 

th 

rough  th(^ 

Bearing  Su 

^C^"-^    " 

fa 

ce. 

(Side 

1&2), 

y^ 

. 

^ 

^^-^ 

y 

.Side 

I 

— 

/ 

^ 

3^- 

^Side 

11 

__ 

i 

y 

^ 

^=^»— 

— 1 

-^^ 

N 

Direction  of 
Rotation  I." 


Direction  of 
Rotation  11." 


0  200  500     1000      1500      2000     2500    3000    3500 
Eevolutions  per  Minute 


Conditions:  Steel  shaft,  3.6  inches  in  diameter;  bearing  I  of  gun  metal  in  one  piece 
10.4  inches  long;  bearing  II  of  white  metal,  split,  good  mineral  oil,  two  rings  immersed 
1.6  inches;  rings  six  inches  in  diameter. 

FIG.    17.      OIL   DELIVERY   CAPACITY    OF   OIL    RINGS   AT    VARIOUS    SPEEDS. 


corners  that  may  catch  and  interfere  with  their  free  motion.     Also  the  passages 
in  which  they  lie  must  be  free  from  projections  upon  which  they  might  catch. 
Many  split  steel  rings  have  been  used  but  are  open  to  the  objection  of  possible 
7 


98 


BEARINGS  AND  THEIR  LUBRICATION 


catching.  Less  commonly  they  have  been  die  cast  in  halves  from  white  metal 
and  the  parts  fastened  together  by  small  screws  through  a  halved  joint.  The 
preferred  ring  is  continuous,  and  made  either  of  a  hard  white  metal  or  brass. 

Artificial  cooling  is  seldom  resorted  to  with  ring  oiling,  as  there  is  no  good 
way  in  which  the  oil  can  be  cooled.  Bearings  have  been  installed  with  a  ser- 
pentine coil  of  pipe  in  the  oil  well  through  which  cold  water  was  circulated,  but 
such  cases  are  rare. 

There  are  no  data  available  as  to  the  amount  of  oil  which  oil  ring  reservoirs 
should  contain.  The  quantity  should  be  great  enough  so  that  all  particles  will 
have  sufficient  time  to  give  up  their  heat  to  the  surrounding  metal  of  the  reser- 


Drill  and  Counter 
'bore  for  Screw.  ^ 


Countersink  sufficient  to  have 
Head  of  Screw  clear  Surface^ 


Prick.P.unch 


^Round  Edges,' 
FIG.   l8. 


Detail  at  Joint 
on  Split  Rings. 


DESIGN   OF   SPLIT   OIL   RING. 


voir  before  they  are  again  brought  in  contact  with  the  shaft.  In  some  instances, 
originally  hot  running  bearings  have  been  made  safer  by  merely  increasing  the 
quantity  of  the  oil. 

Good  practice  places  a  limit  of  4  in.  as  the  greatest  distance  that  an  oil  ring 
can  be  expected  to  successfully  distribute  oil;  that  is,  if  a  bearing  is  8  in.  long, 
one  oil  ring  at  the  center  is  sufficient.  If  the  bearing  is  longer  than  8  in. 
and  shorter  than  16  in.  two  rings  must  be  used,  and  so  on  in  proportion. 

Fig.  18  shows  an  approved  form  of  split  ring. 

Centrifugal  oilers  are  frequently  used  with  step  bearings.  They  consist  of  a 
rotating  vane,  or  series  of  vanes,  that  pump  the  oil  from  a  well  into  which  it 
drains  by  gravity,  through  a  system  of  pipes  to  the  bearings. 

There  is  a  growing  tendency  toward  an  increased  use  of  greases.  These 
are  fed  through  cups  of  two  general  types.  In  one  the  cover  of  the  cup  is 
screwed  down  at  intervals  by  hand  upon  the  contained  grease,  thus  putting  it 


DESIGN  OF  JOURNAL  BEARINGS 


99 


under  pressure  and  forcing  it  into  the  bearings.     In  the  second  type  the  mass 

of  grease  is  under  constant  pressure  from  a  spiral  spring,  or  from  a  weight. 

Bath  lubrication  is  not  very  common.     Here  the  journal  dips  into  a  well  of 


FIG.    ig.       DEVELOPMENT   OF    WHITE-METAL   LINING    SHOWING   THERMOMETER    LOCATIONS. 


FIG.    20.      DISTRIBUTION   OF  TEMPERATURE   IN  A   BEARING — SEE   FIG.    1 9    (Lasche). 


oil,  and  thus  has  an  opportunity  to  take  up  all  that  it  will.    This  is  the  method  of 
lubrication  used  by  Tower  in  his  celebrated  friction  experiments. 

Flooded  lubrication  is  the  supplying  of  oil  to  bearings  by  means  of  a  pump, 
but  at  no  very  great  pressure.     A  typical  example  is  the  system  used  in  oiling 


lOO  BEARINGS  AND  THEIR  LUBRICATION 

the  interior  bearings  of  shafts  and  gears  in  certain  constant  speed  drive 
milling  machines;  also  some  forms  of  steam  turbine  bearings. 

Forced  lubrication  consists  in  supplying  oil  to  bearings  under  pressure  usually 
from  15  to  25  lb.  per  square  inch  for  horizontal  bearings.  As  a  rule  the  oil  is 
cooled  by  artificial  means,  although  this  is  not  inherent  in  the  system.  It  is 
important  that  the  oil  shall  pass  through  the  bearing  at  such  a  rate  of  flow  that 
it  has  time  to  absorb  the  heat  of  friction  and  take  it  away.  This  is  the  reason  for 
introducing  cool  oil  at  a  point  not  far  distant  from  that  of  maximum  pres- 
sure. In  this  way  intense  local  heating  is  avoided.  Fig.  19  is  reproduced 
from  Lasche's  work  and  shows  the  development  of  a  white  metal  lining.  The 
numbered  points  show  where  thermometers  were  brought  very  close  to  the 
inner  surface  of  the  lining,  and  temperature  readings  taken. 

Fig.  20  shows  in  a  solid  diagram  the  observed  temperatures.  The  mini- 
mum temperature  was  at  12,  and  the  maximum  at  about  19,  where  the  pres- 
sure was  greatest. 

OIL  GROOVES 

Oil  grooves  should  be  designed  according  to  the  following  principles: 

A,  The  general  direction  of  grooves  should  be  at  right  angles  to  the  direction  of 
motion.  B,  They  should  stop  short  of  the  ends  of  the  bearing,  to  prevent  leak- 
age. C,  For  conditions  of  considerable  speed  no  groove  must  pass  length- 
wise of  the  bearing  in  the  region  of  greatest  pressure;  even  a  diagonal  or  cir- 
cumferential groove  should  not  be  used  on  the  pressure  side  as  this  will  to 
some  extent  at  least  reduce  the  oil  film  thickness.  It  is  quite  possible  that  for 
very  slow  speed  bearings  grooves  in  the  pressure  area  are  best.  D,  The  edges 
of  all  oil  grooves  and  the  meeting  edges  of  split  boxes  must  be  carefully  rounded 
over. 

Fig.  21  shows  the  oil  grooving  practice  of  the  Westinghouse  Electric  and 
Manufacturing  Company.  The  grooves  vary  from  3/32  to  1/4  in.  in  width 
and  from  3/32  to  1/8  in.  in  depth. 

If  grooves  are  cut  clear  through  to  the  end,  oil  leakage  will  take  place.  At 
the  same  time,  if  a  journal  and  its  bearings  are  very  closely  fitted,  it  may  be 
necessary  to  scratch  or  score  the  surface  from  the  end  of  the  groove  out  in  order 
to  provide  an  air  vent.  The  grooves  at  the  end  of  the  lower  half  of  the  bushings 
are  not  brought  together,  but  are  separated  by  some  little  distance.  This  is  in 
the  region  of  the  greatest  oil  pressure.  If  this  area  is  cut  across  by  oil  grooves, 
it  interferes  with  the  formation  of  a  proper  film.  At  the  meeting  point  of  the 
halves  of  the  bearing  a  large  V-shaped  groove  is  sometimes  cut,  which  serves 
as  a  reservoir  of  oil,  tends  to  relieve  the  side  pressure  on  the  journal,  and  aids 
materially  in  good  lubrication. 


DESIGN  OF  JOURNAL  BEARINGS 


lOI 


At  the  same  time  there  is  a  wide  diversity  of  opinion  in  regard  to  grooves  in 
bearings.  Tradition  says  put  them  in,  and  this  is  usually  done.  Yet  some 
engineers  claim  that  there  should  be  no  grooves  in  the  half  of  the  bearing  that 


One  Ring  Type 


-Leugth  of.  Beariug- 
Waste  Packed 


X 

ting  purposea 

X 

""^ 

i 

,^. 

^ 

Top  of  Bearing 


Omit  Groove  on  this  eud  for  Blct. Bearings 
When  less  than  ?4"oinit  end  Grooved. 
X 


Details  of  Grooves 


Overflow  for 
Automatic  Feed. 


■/         Special  -  Used  when  necessary  to  View  both  Oil  Ringg 
/.  from  common  sight  Hole. 

^         1^ 


igth  oi  J3.eiulug 

ARRANGEMENT   AND    DETAILS    OF    OIL    GROOVES. 


supports  the  load.     Milling  the  box  at  the  split  is  favored  as  this  tends  to  form  a 
small  reservoir. 

Oil  grooves  are  by  no  means  confined  to  the  bearings,  but  are  often  cut  in 
the  journals,  as  shown  by  Fig.  22  from  the  practice  of  the  Step  toe  Shaper  Companjv 
Journals  thus  made  are  run  by  them  in  cast  iron  bearings  with  success.     It  is  the 


H  Oil  Grooves 


't%UA^Il 


FIG.    22.      ARRANGMENT   OF   OIL   GROOVES   IN   A   JONRNAL. 

practice  of  one  English  builder  of  heavy  engines  to  mill  a  flat  spot  lengthwise  of 
the  journals  to  aid  in  their  lubrication.  This  undoubtedly  distributes  the  oil 
over  the  entire  surface  of  the  bearing,  and  while  successful  at  low  speed,  is  prob- 
ably questionable  for  higher  speeds,  as  the  coming  of  the  flat  spot  to  the  region 


I02  BEARINGS  AND  THEIR  LUBRICATION 

of  maximum  pressure  must  tend  to  break  down  the  oil  film  at  each  revolution. 
In  machine  tools  the  Osgood  system  of  oil  grooving  has  been  used  to  a  consider- 
able extent.  It  consists  in  cutting  closed  spirals,  either  single  or  double,  in  the 
lining. 

PROVISIONS  FOR  RADIATING  AND  DISSIPATING  HEAT 

The  heat  radiating  and  dissipating  capacity  of  a  bearing'  must  be  considered 
if  the  engineer  or  designer  is  dealing  with  comparatively  high  pressures  and 
high  velocities.  It  has  been  pointed  out  that  the  customer  usually  determines 
the  maximum  temperature  at  which  a  bearing  may  run,  140°  F.  being  a  common 
limit.  As  we  have  seen,  the  coefficienj;  of  friction  for  all  rough  calculations  may  be 
taken  as  constant  for  the  higher  commercial  speeds.  This  being  so,  the  maximum 
limit  of  speed  is  determined  by  the  amount  of  heat  produced  by  friction  in  a 
unit  of  time,  which  must  not  be  greater  than  the  amount  which  can  be  dealt 
with  by  radiation  and  dissipation  in  the  same  unit  of  time.  Thus  an  increase 
in  the  speed  of  the  shaft  under  these  conditions  is  only  allowable  when  the  excess 
of  frictional  heat  resulting  therefrom  can  be  carried  away  by  air  drafts  or  oil  or 
water  cooling. 

Lasche  conducted  a  long  series  of  experiments  to  determine  the  heat  dissipa- 
tion from  journal  bearings.  From  his  results  Fig.  23  has  been  plotted  with  a 
transformation  of  the  factors  into  English  units.  Three  curves  are  shown.  The 
first  is  for  a  bearing  surrounded  by  a  thin  cast-iron  housing  so  that  the  bearing 
surface  and  the  upper  surface  of  the  housing  can  be  considered  as  at  the  same 
temperature.  In  this  case  it  is  better  to  assume  that  the  radiation  reaches  its 
lowest  limit.  Curve  2  is  for  a  standard  type  of  bearing  in  still  air  and  is  the 
most  useful  of  the  three.  It  can  be  taken  as  representing  an  average  for  the 
radiating  capacity  of  ordinary  bearings  and  used  as  a  basis  from  which  to  calcu- 
late friction  work  and  to  determine  whether  or  not  a  bearing  with  its  natural 
capacity  of  radiation  can  dissipate  the  heat  liberated  by  friction  or  whether 
artificial  cooling  should  be  resorted  to. 

Curve  3  shows  radiation  from  well  ventilated  bearings  with  large  masses  of 
iron  such  as  are  found  in  turbo -generator  construction.  The  unit  work  of 
friction  or  energy  liberated  in  the  form  of  heat,  measured  in  foot  pounds,  is 
given  by  the  formula  fw  V,  in.  which/  is  the  coefficient  of  friction,  w  the  unit 
pressure  in  pounds  per  square  inch  of  projected  area,  and  V  the  rubbing  velocity 
in  feet  per  minute,  or  letting  H  equal  the  rate  of  radiation  per  square  inch  of 
projected  area  expressed  in  foot  pounds  per  minute; 

H=fw  V. 

If  in  the  design  of  a  journal  bearing  the  value  of  the  coefficient  of  friction,/ 
can  be  determined  for  the  particular  working  conditions  of  lubrication,  then 
from  the  above  equation  the  unit  of  radiation  can  be  determined  by  computation 


DfiSIGN  OF  JOURNAL  BEARINGS 


103 


If  we  then  compare  the  unit  thus  found  with  the  radiation  as  traced  from  Fig.  23 
for  the  determined  permissible  rise  in  temperature  of  the  bearing,  it  can  at  once 
be  seen  whether  the  natural  radiation  is  sufficient  to  maintain  a  suitable  tem- 
perature, or  some  artificial  means  of  cooling  must  be  employed. 


c   ^ 

il 


H  m 


S  5 


;:zu 

200 

1 

180 

; 

^ 

^ 

/ 

y 

X 

160 

/ 

/ 

/ 

/ 

/ 

/ 

140 

/ 

/ 

/ 

/  60 
/  ^ 

A 

w 

^ 

120 

1 

/ 

•0 
c 

<^ 

/' 

100 

1 

/i 

Sy 

A 

y 

/>s 

\^ 

^ 

, 

80 

[5 

/  ^ 

J 

# 

4^ 

^ 

■1 

/ 

/ 

^0^ 

^' 

60 

1 

/ 

/ 

^ 

f 

-^ 

if 

/ 

A' 

.'* 

40 

y 

/ 

/' 

/ 

/ 

20 

V 

/ 

/ 

0 

120  240  360  480  600  720  840  960 

Radiation  in  Foot  Pounds  per  Minute  per  Square  Inch  of  Projected  Area  of  Bearing  Surface. 
FIG     23.      CHART  SHOWING  HEAT  RADIATED  PER  UNIT  OF  TIME  FOR  THREE  TYPES  OF  JOURNAL 

BEARINGS. 


Good  judgment  must  be  exercised  in  making  use  of  these  curves,  for  the 
conduction  of  heat  to  a  bearing  along  its  shaft,  as  in  steam  generating  ma- 
chinery, may  play  an  important  part  in  the  amount  of  heat  that  must  be  dissi- 


I04 


BEARINGS  AND  THEIR  LUBRICATION 


pated.  Again,  if  the  bearing  is  surrounded  by  large  masses  of  iron,  with  a 
large  amount  of  exposed  radiating  surface,  the  bearing  will  get  rid  of  more 
heat  than  if  the  surrounding  walls  are  comparatively  thin  and  the  radiating 
surface  comparatively  small     It  is  sometimes  necessary  merely  to  increase  the 

126 


108 


90 


P. 

s 

^ 

H 

g 

fl 

\- 

Ti 

n 

•o 

to 

li 

w 

^ 

> 

?i    54 

rt 

.^ 

g 

O 

U 

h 

bo 

cS 

^ 

Pm 

u 

S  20 


18 


70 

^ 

j^^ 

^^ 

120 

0— 

^ 

^ 

^ 

^ 

60 

y 

y 

^ 

1000 

y 

/ 

^ 

^ 

/ 

/ 

^ 

/ 

/ 

^ 

800 

/ 

/ 

/ 

y' 

^ 

^ 

^ 

/ 

/ 

y 

^ 

/ 

/ 

) 

/ 

^ 

/ 

/ 

/ 

/ 

^ 

y 

600 

At\ 

/ 

/ 

/ 

/ 

y 

^^ 

^- 

-^ 

/ 

/ 

/ 

y 

^ 

-^ 

^ 

/ 

/ 

/ 

/ 

y 

\^ 

r 

/ 

/ 

^ 

y 

30 

/ 

y 

— 

, 

400 

/ 

/ 

y 

^ 

/ 

/ 

^- 

?C^ 

^t^ 

Uxi- 

/ 

^ 

'v^^ 

r^^ 

?n 

/ 

^ 

^ 

^^^ 

^ 

y' 

/ 

/ 

10 

0 

10 


90 


100 


!0  30  40  50  60  70  80 

Pressure  Lbs.  per  Sq.  In.  Proj.  Area. 
FIG.    24.       TEMPERATURE    RISE    OF   OIL   RING    BEARING    IN    STILL   AIR— ROOM   TEM- 
PERATURE   25    DEGREES   CENTIGRADE,    77    DEGREES   FAHRENHEIT. 

size  of  the  oil  reservoir.  Again,  an  increase  in  the  extent  of  the  surface  of  a 
bearing  housing  may  be  all  that  is  required  to  cause  a  bearing  to  run  within  its 
permissible  rise  in  temperature.  Another  factor  is  the  proximity  of  rotating 
engine,  generator,  or  motor  parts,  which  set  up  windage  and  cause  a  current 


DESIGN  OF  JOURNAL  BEARINGS 


105 


of  air  to  pass  over  the  bearing.  All  of  these  factors,  as  well  as  any  others  which 
directly  affect  the  radiation  or  dissipation  of  heat  must  be  taken  into  account 
in  determining  the  temperature  at  which  a  given  bearing  will  run  in  service. 
The  underlying  principle  may  be  stated  thus:  The  work  of  friction  per  unit  of 
bearing  area  must  not  be  greater  than  the  rate  of  radiation  for  the  same  unit 
area,  and  for  the  determined  working  temperature. 

Lasche^  developed  an  equation  expressing  the  relationship  of  the  coefficient 
of  friction,  unit  pressure,  and  rise  in  temperature,  which  transformed  into 

eo 


72 


-o  54  « 


36 


bo 

P 
U 

^    18 

en 


1 

120( 

Ft 

40 

^ 

100( 

Ft 

^ 

•^ 

30 

^ 

y 

800 

Ft. 

y 

y 

^ 

^ 

^ 

^ 

/ 

^ 

y 

600 

Ft. 

y 

^ 

-^ 

^ 

20 

/ 

-X 

y 

-^ 

X 

^ 

400 

Ft. 

/ 

^ 

^ 

^^ 

7^ 

n. 

,— - 

y 

^ 

rsv 

bs 

f5 

Tyt 

10 

^ 

^ 

-^ 

^^ 

^ 

r 

10 


20 


90 


100 


0  40  50  60  70  8 

Pressure.  Lbs.  per  Sq.  In.  Proj.  Area. 

FIG.    25.      TEMPERATURE    RISE    OF   OIL   RING    BEARING    FOR    WELL   VENTILATED    CONDITION 

ROOM  TEMPERATURE   25    DEGREES    CENTIGRADE,    77    DEGREES    FAHRENHEIT. 


fw= 


English  unit  is: /w(/— 32)  =  51.2,  in  which /is  the  coefficient  of  friction,  w  the 
pressure  in  pounds  per  square  inch,  and  /  the  final  bearing  temperature. 
By  transformation 

/-32 

In  the  section  dealing  with  values  of  the  coefficient  of  friction  it  was  shown  that 
for  conditions  wherein  a  perfect  oil  film  can  be  formed  the  f rictional  resistance  for 
a  given  velocity  is  practically  a  constant  and  independent  of  the  pressure.    Thus, 

iPor  a  valuable  discussion  of  this  equation  see  Elements  of  Machine  Design,  Kimball  and  Ban, 
page  247. 


lo6  BEARINGS  AND  THEIR  LUBRICATION 

as  far  as  ordinary  designing  is  concerned,  we  may  assume  that  the  coefficient  of 
friction  for  perfect  lubrication  and  all  velocities  greater  than  500  ft.  per  minute  is 
a  constant,  and  the  equation  given  above  may  be  used  without  serious  error. 

By  adding  the  factor  V  to  each  side  the  right  hand  member  becomes  equal 
to  the  frictional  loss  per  unit  of  projected  area  expressed  in  foot  pounds  per 
minute,  as  below: 

fwV=     

t-32 

This  equation  can  now  be  used  to  compute  the  heat  liberated  by  friction 
in  a  perfectly  lubricated  high  speed  journal  bearing  for  a  given  velocity  in  feet 
per  minute  at  a  given  working  temperature  in  degrees  Fahrenheit. 

This  gives  us  a  unit  of  radiation  to  compare  with  the  curves  of  Fig.  23  to 
determine  whether  or  not  artificial  cooling  must  be  used. 

The  General  Electric  Company  has  carried  on  an  extensive  series  of  tests 
to  determine  the  rise  in  temperature  for  bearings  of  a  type  used  by  them  in 
electrical  machinery  and  under  various  conditions;  the  results  of  these  experi- 
ments are  given  in  the  curves  of  Figs.  24,  25,  and  26.  The  first  is  for  a  ring 
oiling  bearing  running  in  still  air,  as  in  a  tightly  closed  room  with  an  air  tempera- 
ture of  77°  F.  Fig.  25  is  for  the  same  kind  of  bearing  well  ventilated.  Fig.  26 
is  for  a  bearing  running  with  forced  lubrication  and  three  different  rates  of  oil 
feed.  In  each  case  the  lubricating  oil  was  what  is  known  as  ''dynamo  and 
shafting  oil,"  having  a  specific  heat  of  0.4,  a  weight  per  gallon  of  7.27  lb.,  and 
a  viscosity  at  70°  F.  of  80.3  Doolittle.  Curves  24  and  25  are  for  a  range  of 
speed  from  400  to  1200  ft.  per  minute.  The  curves  of  Fig.  26  are  for  rubbing 
speeds  up  to  3,000  ft.  per  minute  and  for  four  different  pressures  per  square  inch 
of  projected  area,  namely,  35,  50,  75,  and  100  lb.  For  pressures  and  speeds 
beyond  the  limit  of  this  chart  it  is  wise  to  resort  to  water  jacket  cooling  to 
remove  the  liberated  heat. 

RULES  FOR  BEARING  DESIGNS 

This  section  on  bearing  design  can  be  closed  in  no  better  way  than  to  give 

the  rules  for  bearing  designs  as  adopted  by  the  Committee  of  Mechanical  Design 

of  the  General  Electric  Company,  issued  in  connection  with  the  preceding  charts. 

These  rules  are  applicable  to  the  design  of  the  great  majority  of  generator 

and  motor  bearings. 

The  temperature  of  any  bearing  having  a  given  radiating  surface  is  deter- 
mined by  the  amount  of  work  lost  by  friction.  This  work  (in  foot  pounds  per 
minute)  =/  W  V,  where, 

/=  coefficient  of  friction, 
W=\o2id  on  the  bearing  in  pounds, 
V=  surface  speed  in  feet  per  minute. 


DESIGN  OF  JOURNAL  BEARINGS 


107 


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Io8  BEARINGS  AND  THEIR  LUBRICATION 

Assuming  that  the  load  remains  fixed  it  is  evident  that  the  work  lost  and 
therefore  the  amount  of  heat  generated  becomes  smaller  as/  and  V  become 
smaller;  or,  in  other  words  the  smaller  the  bearing  diameter  for  a  given  load  the 
less  it  will  heat.  Another  advantage  of  small  diameter  is  that  the  coefficient 
of  friction  decreases  with  the  surface  speed.  It  is,  therefore,  very  desirable 
in  laying  out  bearings  to  keep  the  diameter  as  small  as  possible,  consistent  with 
sufficient  strength  of  shaft  and  suitable  deflection  of  the  journals  both  inside 
and  outside  the  bearings.  It  is  also  very  desirable  to  so  dimension  bearings 
that  they  are  fairly  well  loaded,  in  order  to  avoid  bulky  machines  and  also 
because  the  coefficient  of  friction  rises  quite  rapidly  when  the  load  is  less  than 
50  lb.  per  square  inch  of  projected  area. 

When  calculating  the  projected  area  of  any  bearing,  especially  if  it  is  to  be 
heavily  loaded,  the  amount  of  space  lost  through  the  drain  grooves  at  both  ends 
must  be  deducted.  This  is  particularly  important  when  the  length  of  the  bear- 
ing is  small  in  proportion  to  the  diameter. 

It  is  also  necessary — and  this  applies  to  all  forms  of  lubrication — that  there 
be  no  sharp  corners  on  the  edges  of  the  oil  distributing  grooves  or  channels,  but 
that  these  be  gradually  cased  off  so  that  the  oil  can  be  drawn  in  between  the 
journal  and  the  bearing.  Sharp  corners  are  invariably  oil  wipers  and  often 
absolutely  prevent  proper  lubrication. 

The  heat  generated  in  any  bearing  may  be  dissipated: 

1.  By  radiation  from  the  housings  and  conduction  by  the  shaft. 

2.  By  forcing  cooled  oil  through  the  bearing. 

3.  By  surrrounding  the  bearing  by  some  form  of  water  jacket. 

Other  forms  of  artificial  cooling  are  possible  but  these  described  under  2 
and  3  are  the  most  commonly  used. 

I. — BEARINGS    WITHOUT   ARTIFICIAL    COOLING 

Bearings  of  this  type  are  usually  lubricated  by  oil  rings  or  similar  devices 
or  by  gravity  feed.  It  is  essential  that  an  abundant  supply  of  oil  be  delivered 
to  all  parts  of  the  bearing  by  suitably  arranging  the  channels  so  that  a  perfect 
film  will  be  maintained  at  all  times  between  the  journal  and  bearing,  and  that 
there  is  no  opportunity  for  the  oil  forming  this  film  to  escape  through  openings 
or  grooves  at  the  points  of  greatest  pressure  and  thus  allow  the  metals  t6  come 
in  contact. 

The  heat  generated  in  bearings  having  no  artificial  cooling  is  conducted 
away  and  radiated  by  the  housings.  The  great  variation  in  the  design  of 
bearing  housings  and  the  different  conditions  of  ventilation,  etc.,  make  it 
extremely  difficult  to  predetermine  the  ultimate  temperature  of  such  bearings 
with  any  great  accuracy,  and  it  is  always  necessary  to  allow  a  considerable 
margin  of  safety. 


DESIGN  OF  JOURNAL  BEARINGS  109 

Fig.  1 1  covers  the  range  of  pressure  and  speed  ordinarily  permissible  in  this 
type  of  bearing,  while  Figs.  24  and  25  show  the  ultimate  temperatures  for  differ- 
ent speeds  and  loads.  These  curves  were  made  up  from  the  readings  obtained 
from  special  bearings  and  afterward  checked  by  the  test  records  of  a  large 
number  of  machines — both  of  the  pillow  block  and  shield  types — which  have 
gone  through  the  testing  department  during  the  last  few  years.  Fig.  24  shows 
the  temperatures  to  be  expected  under  the  most  unfavorable  conditions,  that  is, 
of  a  bearing  so  situated  that  no  current  of  air  can  circulate  about  it,  and  there- 
fore cooled  by  radiation  only.  There  is,  however,  a  considerable  circulation  of 
air  about  most  machines,  due  to  the  fanning  action  of  the  revolving  parts,  and 
the  ultimate  temperatures  to  be  expected  in  such  cases  are  shown  on  the  curves 
of  Fig.  25.  These  curves  apply  to  the  great  majority  of  open  generators  and 
motors,  both  of  the  pillow  block  and  end  shield  types.  When  the  machine  is 
enclosed,  or  the  free  circulation  of  air  in  any  way  interrupted,  higher  tempera- 
tures will  result,  until  finally  the  conditions  of  Fig.  24  are  reached. 

A  part  of  the  heat  of  the  bearings  of  motors  and  generators  is  usually  con- 
ducted away  by  the  shaft  and  radiated  by  the  spider  and  other  revolving  parts. 
When  machines  are  totally  enclosed  or  are  connected  to  other  machinery  whose 
temperature  is  high,  heat  may  be  transmitted  through  the  shaft  to  the  bearing, 
thus  raising  the  latter's  temperature,  and  due  allowance  must  be  made  for  this. 

2. — COOLING  BY  FORCED  LUBRICATION 

When  bearing  pressures  and  speeds  are  unusually  high  it  is  often  necessary 
to  force  oil  under  pressure  into  the  bearings  and  advantage  is  often  taken  of 
this  to  keep  the  heating  down  by  artificially  cooling  the  oil. 

The  temperature  of  bearings  is  also  sometimes  kept  down  by  forcing  cooled 
oil  through  them,  when  otherwise  forced  lubrication  is  not  necessary.  This 
method,  although  used  to  a  very  considerable  extent,  is  usually  not  as  efficient 
as  a  water  jacket. 

For  all  practical  purposes,  it  may  be  considered  that  the  entire  heat  generated 
is  taken  away  by  the  oil,  and  it  is  therefore  possible  to  predetermine  the  bearing 
temperature  with  considerable  accuracy.  Fig.  26  shows  the  ultimate  temper- 
ature of  bearings  using  the  quantities  of  oil  most  commonly  pumped  through, 
and  with  the  assistance  of  these  curves  the  necessary  amount  can  be  determined. 
Intermediate  speeds  and  pressures  can  be  easily  interpolated. 

For  pressures  and  speeds  beyond  the  limits  of  this  table,  it  is  advisable  to 
resort  to  water  jacket  cooling. 

In  arranging  bearings  for  this  form  of  lubrication,  care  must  be  taken  to 
force  the  oil  to  the  point  where  the  work  is  being  done,  as  otherwise  the  oil 
coming  from  the  bearing  may  be  comparatively  cool,  while  the  bearing  itself 


no  BEARINGS  AND  THEIR  LUBRICATION 

is  much  too  warm.  In  addition  to  this,  it  means  that  an  excessive  amount  of 
oil  must  be  pumped  through  the  bearing  requiring  unnecessarily  large  pumps, 
pipings,  etc.  Experiments  with  such  bearings  show  that  when  oil  begins  to 
run  out  of  the  ends  quite  freely,  nothing  is  gained  by  forcing  through  a  larger 
quantity. 

3. — COOLING  BY  WATER  JACKET 

A  properly  designed  water  jacket  will  carry  away  a  very  much  larger  amount 
of  heat  than  will  any  form  of  forced  oil  lubrication,  as  the  specific  heat  of 
water  is  very  much  higher.  As  is  the  case  of  forced  oil  lubrication  the  ultimate 
temperature  of  a  well  designed  water  jacket  bearing  can  be  very  accurately 
determined.  It  is  essential  that  the  pipes  or  channels  be  located  as  close  as 
possible  to  the  surface  of  the  lining  where  the  work  is  being  done,  in  order  that 
the  heat  generated  may  be  absorbed  without  danger  of  damage  to  the  lining. 
Water  circulated  at  some  distance  away  from  the  lining  surface  is  of  compara- 
tively little  assistance,  as  heat  may  be  generated  so  rapidly  that  the  lining  will 
be  destroyed  before  the  heat  strikes  the  jacket.  The  water  passages  must  also 
be  so  arranged  that  an  even  and  continuous  circulation  is  kept  up  in  all  parts. 

With  properly  constructed  passages,  it  is  safe  to  assume  that  heat  may  be 
removed  at  the  rate  of  from  3  to  5  horse  power  for  each  gallon  of  water  circulated 
per  minute,  and  where  the  conditions  are  unusually  good  and  the  jacketing 
carefully  arranged  from  10  to  12  horsepower  per  gallon  can  be  dissipated. 

With  water  jackets  any  suitable  method  of  lubrication  may  be  used  which 
will  insure  at  all  times  a  good  film  of  oil  between  journal  and  bearing. 


Lubricating  oils. 


SECTION  VI 

Lubricants  for  Bearings 

The  subject  of  lubricants  for  bearings  is  one  of  great  complexity.  The 
chemical  analysis  of  oils  and  greases  is  very  difficult  and  is  work  that  can  be 
successfully  carried  on  only  by  a  specially  trained  chemist.  Many  oils  are 
mixed  or  blended  and  for  the  user  there  are  no  simple  tests  by  means  of  which 
he  can  determine  just  what  the  constituents  of  a  particular  kind  of  oil  are. 
At  the  best  he  can  only  make  a  few  simple  physical  tests  and  then  study  the 
behavior  of  the  oil  in  use. 

The  following  paragraphs  will  not  touch  upon  the  chemical  side,  but  merely 
point  out  some  of  the  important  kinds  of  oils  and  greases,  their  applications, 
and  some  of  the  physical  characteristics  that  they  should  have. 

CLASSIFICATION  OF  LUBRICANTS 

All  lubricants  fall  into  three  classes:  oils,  greases,  and  solids. 

High  pressure  cylinder  oil, 

Low  pressure  cylinder  oil, 

Engine  oil, 

Dynamo  and  shafting  oil, 

Machinery ^oil  for  heavy  duty, 

Machinery  oil  for  light  duty, 

Loom  oil, 

Spindle  oil. 

Non-fluid  oil. 

Soft  grease. 

Hard  grease. 

f  Graphite, 
Solid  lubricants <  Sulphur, 

[  Talc. 

The  oils  are  again  divided  into  three  divisions  as  regards  origin,  mineral, 
vegetable  and  animal.  The  mineral  oils  are  petroleum  products  and  in  this 
country  are  spoken  of  as  of  two  kinds,  oils  with  a  paraffine  base  and  oils  with 
an  asphalt  base ,  the  Pennsylvania  product  being  representative  of  the  first  class 
and  the  Texas  product  of  the  second.  Oils  from  certain  localities  have  come 
to  be  recognized  as  standards  for  some  purposes,  as,  for  instance,  the  best  cyl- 
inder oils  are  supposed  to  be  made  from  Franklin  stock. 

Among  vegetable  oils,  colza,  rape  seed,  olive,  and  palm  oils  are  good  lubri- 
cants. Castor  oil  has  good  lubricating  qualities  but  oxidizes  and  gums. 
Among  animal  oils,  lard,  tallow,  sperm,  and  whale  are  all  used  both  pure  and 
for  compounding  with  mineral  stock. 

Ill 


Lubricating  greases. 


112 


BEARINGS  AND  THEIR  LUBRICATION 


Graphite  is  the  only  solid  lubricant  that  is  used  to  any  extent,  and  it  is  recog- 
nized as  one  of  our  most  valuable  lubricants.  It  is  employed  in  a  solid  form 
fitted  into  recesses  in  bearing  linings,  in  powdered  form  mixed  with  oil  and  in 
the  form  of  deflocculated  graphite  mixed  with  oil  or  water.  This  last  form  is  a 
special  product  of  the  Acheson  Graphite  Company.  Sulphur  was  formerly 
employed  for  emergency  use  on  over-heated  bearings,  prior  to  the  common  use  of 
graphite.  Ground  talc  or  soapstone  free  from  grit  has  had  a  limited  use  for 
the  same  purposes. 

Water  has  been  used  as  a  lubricant,  particularly  for  the  steps  of  vertical 
water  turbines  and  vertical  steam  turbines.  Its  action,  however,  is  not  thai  of 
a  true  lubricant.  In  the  vertical  steam  turbine  and  some  water  turbines  it  serves 
as  a  means  to  separate  the  plates  of  the  steps.  It  is  pumped  in  under  pressure 
and  floats  the  load.  In  other  water  turbines  having  wood  block  and  metal  cup 
bearings  it  removes  the  heat  liberated  by  friction  and  prevents  the  bearing  from 
burning  and  cutting. 

Many  other  kinds  of  oils  are  used  but  they  are  often  looked  upon  as  adul- 
terants; for  instance,  cotton-seed  oil  is  frequently  mixed  with  lard  oil.  As 
mentioned  above,  their  detection  is  very  difficult  and  should  only  be  undertaken 
by  the  trained  chemist.  Just  what  oils  may  be  used  as  adulterants  at  a  given 
time  depends  largely  upon  the  relative  market  values  of  the  oils. 

Turning  to  physical  tests,  the  oil  user  is  most  largely  interested  in  the 
qualities  of  viscosity  and  body.  The  viscosity  of  an  oil  is  the  property  that 
determines  the  rate  of  free  flow. 

The  property,  body,  is  difficult  to  define.  The  user  gets  his  idea  of  it  by 
dipping  his  thumb  and  forefinger  into  the  oil  and  rubbing  them  together.  The 
degree  of  oiliness  or  greasiness  is  an  indication  to  him  of  its  body.  It  has  been 
defined  as  that  property  of  an  oil  that  influences  the  change  in  viscosity  when 
the  oil  is  under  pressure;  or  as  that  property  that  influences  the  intensifying  of 
the  viscosity  in  that  portion  of  the  oil  film  within  the  region  of  attraction  of  the 
surface  molecules  of  the  metals  of  bearing  and  journal. 

Professor  Kingsbury,  Trans.  A,S.  M.  E.,  1903,  page  147,  gives  these  rela- 
tions of  viscosity  and  body. 


With  increase  of 


Where  the  viscosity  is  effective, 
the  coefficient  of  friction 


Where  the  body  is  effective,  the 
coeflicient  of  friction 


Pressure. .  .    . 

Speed 

Temperature. 

Viscosity 

Body 


Decreases. 
Increases . 
Decreases. 
Increases . 


Increases. 
Decreases. 
Increases, 
Decreases. 
Decreases. 


LUBRICANTS 


113 


From  this  it  is  seen  that  the  effects  of  body  and  viscosity  are  in  nearly  all 
respects  diametrically  opposite  and  it  must  necessarily  be  very  difficult  to 
derive  reliable  information  regarding  the  lubricating  values  of  oil  from  friction 
tests  in  which  the  effects  of  viscosity  and  body  are  not  separately  recognizable. 
Thus  we  have  but  very  litde  information  that  is  of  value  on  the  body  of  oils. 
As  a  general  rule,  mineral  oils  have  much  less  body  than  animal  and  vegetable 
oils.  Thus  of  a  mineral  oil  and  lard  oil  of  approximately  the  same  viscosity 
the  lard  oil  would  have  decidedly  the  greater  body.  On  the  other  hand,  some 
cylinder  oils  of  mineral  origin  greatly  exceeding  lard  oil  in  viscosity  will  also 
have  a  greater  body. 

It  is  not  uncommon  to  mix  a  small  percentage  of  animal  oil  with  mineral  oil 
in  order  to  increase  the  body.  For  illustration,  turn  to  the  latter  part  of  this 
section  where  specifications  for  oils  used  by  the  United  States  War  and  Navy 
Departments  are  given. 

During  the  past  few  years  there  has  been  a  tendency  to  use  more  and  more 
grease  for  the  lubrication  of  machinery  bearings.  In  comparing  the  relative 
qualities  of  a  grease  and  an  oil,  the  former  must  be  tested  at  the  running  tem- 
perature of  the  bearing  where  it  is  to  be  used.  In  general  the  lower  the  speed 
the  thicker  should  be  the  lubricant.  With  heavy  pressures  and  low  speed  it  is 
difficult  to  form  a  film  of  oil  and  maintain  it.  This  is  specially  true  when  the 
machine  is  started  after  being  shut  down.  While  idle  the  oil  is  squeezed  out 
from  between  the  journal  and  bearing  surfaces  and  there  is  metallic  contact 
which  produces  abrasion  as  soon  as  motion  begins  with  an  attending  rise  in 
temperature  and  thinning  of  the  oil  if  that  is  used.  For  such  work  greases  are 
especially  suitable,  as  the  solid  or  semisolid  substance  is  not  squeezed  out  from 
the  journal  in  the  way  a  liquid  is  acted  upon. 

Professor  Goodman,  Proc.  Inst.  C.  £.,  Vol.  LXXXIX,  page  432,  gives  com- 
parative results  of  three  oils  and  one  grease.  The  nominal  pressure  per  square 
inch  was  85  lb. 


Lubricants 


Natural  coefficient  of 
friction 


Coefficient  of  friction  after 
standing  16  hours 


Machinery  oil 

Valve  oil J  A 

Thick  and  viscous (^  B 

Hard  solid  grease 


o . 0084 
0.0252 
0.0329 
0.0350 


0.192 
0.147 
0.1715 
o .  0(;o 


It  will  be  noted  that  the  coefficient  of  friction  after  standing  is  only  about 
half  as  great  for  the  grease  as  for  the  oils. 


114 


BEARINGS  AND  THEIR  LUBRICATION 


On  the  other  hand,  a  grease  should  not  be  used  unless  the  conditions  for 
feeding  it  to  the  bearing  are  such  that  it  will  reach  all  parts  of  the  rubbing 
surfaces.  An  incident  given  by  W.  H.  Booth,  American  Machinist^  Vol.  XXXII 
part  2,  page  790,  bears  on  this  point. 

*' A  set  of  pumps  driven  by  a  worm  and  wheel  were  required  to  be  tested  for 
general  sufficiency.  The  worm  was  below  the  wheel  and  there  was  a  thrust 
bearing  in  a  narrow  cylindrical  recess  out  of  the  oil  casing  which  inclosed  the 
worm  and  wheel.  A  thick  oil  was  filled  into  the  casing  and  the  test  was  made. 
The  thrust  bearing  soon  grew  unbearably  hot,  and  it  looked  as  if  the  pump 
would  be  condemned.  Quite  suddenly  the  box  went  comparatively  cool  and 
no  further  heating  took  place.  The  cooling  coincided  with  the  thinning  or 
liquefying  of  the  oil.  At  first  this  was  so  very  thick  that  it  could  not  freely 
enter  the  thrust  chamber  against  the  tendency  of  the  worm  to  draw  the  oil 
away  from  the  thrust  recess  and  pile  it  up  on  the  opposite  side  of  the  casing, 
but  this  piling  up  ceased  when  the  oil  became  hot  and  thin,  for  it  could  then 
get  well  into  the  thrust  collars,  and  these  cooled  at  once,  and  the  test  was 
declared  good.  A  thinner  oil  would  have  run  from  the  start  in  all  likelihood, 
and  this  is  confirmed  by  the  experience  of  many,  that  a  thin  oil  is  a  better 
lubricant  than  a  thick  oil  up  to  the  point  where  the  pressure  begins  to  squeeze 
the  oil  from  between  the  surfaces. " 

From  Tower's  experiments,  Proc.  Inst.  M.  E.,  1883,  the  following  comparison 
of  five  oils  and  one  grease  is  taken: 


Lubricant 


Mean  resistance, 
pounds 


Sperm  oil 

Rape  oil 

Mineral  oil 

Lard  oil 

Olive  oil 

Mineral  grease 


Per  cent 


100 
106 
129 

135 
135 
217 


The  second  column  gives  the  actual  frictional  resistance  at  the  surface  of 
the  journal  per  square  inch  of  bearing  area  at  a  speed  of  300  revolutions  per 
minute,  and  for  nominal  loads  from  100  to  310  lb.  per  square  inch.  The 
percentage  figures  may  be  taken  to  represent  the  relative  thickness  or  body  of 
the  various  oils,  and  in  their  order,  if  not  in  numerical  proportion,  the  relative 
weight  carrying  ability.  Thus  it  is  seen  that  the  sperm  oil  has  by  far  the 
greatest  lubricating  value  and  the  least  weight  carrying  ability. 

It  must  be  recognized  that  these  values,  however,  hold  only  for  the  conditions 


LUBRICANTS  I15 

of  investigation.  For  greater  pressures  or  higher  temperatures  the  relative 
order  of  excellence  might  easily  be  changed  with  some  of  the  heavier  oils  leading. 

Another  very  important  quality  of  any  lubricating  oil  is  that  it  shall  be  free 
from  acids  and  alkalies.  Animal  and  vegetable  oils  must  be  watched  more 
closely  in  this  particular  than  mineral  oils.  As  a  rule  mineral  oils  have  no 
action  on  metals  unless  they  contain  sulphur,  and  then  there  is  no  liability  of 
harm  except  at  high  temperatures.  Among  the  organic  oils  lard  oil  and  pure 
animal  oils  in  general  have  little  action;  on  the  other  hand,  olive  and  vegetable 
oils  generally  produce  corrosion. 

Other  points  to  be  investigated  in  lubricating  oils  are  the  flash  and  fire 
points,  the  amount  of  contained  water,  the  amount  of  tarry  ancjl  suspended 
matter,  the  volatility  and  action  under  saponification.  For  good  practice  in 
regard  to  these  points  see  the  United  States  War  Department  specifications 
in  the  latter  part  of  this  section. 

QUANTITIES  OF  OIL  USED 

As  is  to  be  expected,  there  are  few  data  available  for  the  quantity  of  oil 
used  for  bearing  lubrication.  The  variation  in  type  of  bearing,  rubbing  speed, 
pressures,  quality  of  the  oil  and  its  method  of  application,  all  effect  this  question. 

As  representative  of  American  practice  in  the  lubrication  of  steam  engine 
cylinders.  Table  35  is  reproduced  from  Power,  Feb.  15,  1910,  page  303.  A 
glance  at  the  last  column  shows  that  the  average  is  about  one  pint  for  every 
million  square  feet  of  cylinder  surface  rubbed  over. 

For  bearings  lubricated  by  forced  pressure  and  water  cooled  1/50  gal.  per 
square  inch  of  projected  area  per  minute  is  good  practice. 

For  bearings  with  flooded  lubrication  and  cooling  of  the  oil  i/io  gal.  per 
square  inch  of  projected  area  per  minute  is  frequent  practice. 

For  the  step  bearings  and  guide  bearings  of  Curtis  vertical  steam  turbines 
Table  36  can  be  considered  as  presenting  average  practice. 


ii6 


BEARINGS  AND  THEIR  LUBRICATION 


29 
30 
31 
32 
33 
34 

35 
36 
37 
38 
39 
40 

41 
42 

43  I 

44 

45 

46 

47 

48 

49 

50 


Esti- 
mated 
h.  p. 


1,350 
67s 
67s 
975 
975 
650 
650 

575 
875 
450 
650 
290 
290 
200 
200 
400 
450 
290 
200 

225 
225 
225 
450 
200 
180 
160 
160 
70 
450 

4SO 
290 
290 
290 
200 
160 
290 
200 
130 
250 
160 
160 
130 
130 
ISO 
130 
120 
160 
130 
130 
100 


Description  of  engine 


Cylin- 
ders 


Stroke 


I 
Corliss  triple  expansion ;  20-34-52  i 


Cross  compound. 


I 


Cross  comp.  St.  Louis  Corliss. 

Corliss  compound 

Hamilton  Corliss  cross  comp. 
Hamilton  Corliss  cross  comp  . 


Corliss  compound 

Corliss  tandem  compound. 
Russell  cross  compound.  .  . 
Cross  comp.  condensing.  .  . 

>  Cross  compound 


Harris  standard  cross  comp  . . 
Harris  standard  cross  comp  .  . 
Cross  comp.  St.  Louis  Corliss. 

Corliss  cross  comp 

Phoenix  Iron  Works 

Russel  tandem  comp 


Corliss  cross  compound. 


Buckeye  tandem  comp 

IngersoU  cross  compound 

American  ball  compound 

Westinghouse  automatic  comp. 

Cross  comp.  Meyer  valves 

Tandem  comp.  piston  valve.  .  . 

IngersoU  cross  comp 

Piston  valve 

Double  eccentric  corliss 

Corliss 

Harris  Corliss 


Greene 

Sioux  Corliss 

Harris  Corliss 

Corliss 

New  Brown 

Atlas  single  valve  automatic.  .  . 

Ames 

Corliss 

High  speed  automatic 

High  speed  automatic 

Piston  valve  engine 

Slide  valve  throttling  governor. 
Single  valve  auto.  Fitchburg  . . 

Atlas  four  valve .  , 

Harris  Corliss 


h.  p.  26 
1.  p.  52 
22-44 
24-44 
20—36 
20-36 

18-34 
24-42 
16-30 
18-36 

h.  p.  18 
1-  p.  34 

h.  p.  16 
1.  p.  28 
14-28 
18-30 
14-24 

I3-20i 

h.  p.  16 

1.  p.  30 

12-21 

18-30 

102-20^- 

11-19 

11-18 

iof-i8  * 

8-12 

30 

30 

24 

24 

24 

20 

18 

24 

20 

16 

22^ 

18 

18 

16 

16 

I7i 

16 

15^ 

18 

16 

16 

14 


60 
48 
48 
48 
60 
44 
44 

48 
48 
24 
42 
36 
36 
18 
18 
42 
36 
18 
20 
36 
36 
21 
24 
12 
1 1 
14 
18 
12 
48 
48 
48 
48 
42 
44 
42 
48 
36 
36 
30 
16 
36 
16 
16 
16 
30 
14 
15 
36 
38 


Revolu- 
tions per 
minute 


65 


81 

85 

6S 

100 

100 

85 

6  5 

150 

72 

86 

86 

200 

200 

8S 

84 


84 

84 

200 

100 

28s 

300 

16s 

IIS 

150 

95 

84 


100 

105 

6S 

102 

125 

los 
230 
100 
250 
250 
220 
126 
27S 
220 
100 
92 
280 


Steam 
pressure 


140 


140  100  sup 


IIS 
115 


ISO 
150 


50 
no 
no 


IIS 

115 

80 

95 


125 
65 


.  Sq.  ft. 

rubbed 

over  per 

hour 


1,084,000 
265,800 
531,600 
706,500 
694,500 
644,400 
644,400 

555,900 
539,600, 
433,3CO 
428,100 
146,000 
276,000 
150,800 
263,900 
393, coo 
380,000 
376,100 
368,200 
126,800 
237,500 
363,100 
302,000 
277,600 
259.300 
175.400 
172,100 
94,3°° 
357.800 
318,000 
308,100 
302,000 
245,500 
230,200 
207,900 
196,200 
196,200 
188, 8co 
183,500 
173,700 
169,800 
167,800 
167,800 
161,500 
158,200 
156,200 
iSS,9co 
150,900 
146,600 
143,200 


Table  35. — Steam  Engine  Cylinder  Lubrication  (Power) 


LUBRICANTS 


117 


No. 


Oil  used 


Name  or  description 


Price  per 
gal. 


Amount 
used  per 
hr.,  pints 


_     .  I   Pmts  per 

Cost  per.  I      ^.      ^    , 
,  estimated 

hr., 

.  1000  h.  p. 

cents  , 

hours 


Cents  per  1    Pints  per 
estimated  |      million 
1000  h.  p.  I  sq.  ft.  rub- 
hours      I    bed  over 


12 
I? 
14 
15 
16 
17 
18 
19 
20 
21 
22 
23 
24 
25 
26 
27 
28 
29 
30 
31 
32 
33 
34 
35 
36 
37 
38 
39 
40 
41 
42 
43 
44 
45 
46 
47 
48 
49 
SO 


No.  725  cyl.  comp 

Capitol  oil 

No.  650  dark  valve  cyl, 


"600  W." 

Oil  of  beeswax;  600  fire  test  1 
cyl. stock  and  acidless  tallow  J 

Heavy  body  comp 

"600  W." 

Improved  high  pressure  cyl .  .  . 

Best  grade  cyl 

Best  grade  cyl 

Harris  h.  p.  valve 

Rarus 

No.  650  dark  valve 

"600  W." 


"600  W." 

Best  grade  cyl. 
Best  grade  cyl. 

"600  W." 

"600  W." 


No.  725  cyl.  comp 

Best  high  grade  mineral. 

High  grade  mineral 

"600  W." 

"Cyl.  oil  No.  10." 

"Cyl.  oil  No.  10." 


600  W. 


"Cyl.  oil  No.  10." 

Harris  high-press  valve. 

Buckeye  cyl 

Capital  oil 


High  grade  mineral . 
High  grade  mineral. 
Light  colored  oil ... . 
Capitol 


.28 
.75 
•35 
.50 
.28 
.60 

90 

.28 
.40 
.60 
•75 
.60 
.60 
.60 


0.69 
0.466 
0.666 
o.  104 
0.621 
o.  21 

0.25s 

0.46s 

0.2 

0.25 

O  .  1 1  I 

0.64s 
0.323 

o.  184 
o.  131 
0.09 

0.417 
0.167 
o.  167 

0.64s 
0.323 
0.2s 

o.  167 

0.273 
0.083 
0.012 

0.037 
0.C83 
0.417 
0.5 

o.cgi 
0.091 
0.369 
0.91 
o.  125 
o.  19 
o.  167 


O.  22 


Rarus  cyl 

Harris  std.  grade. 
Compounded  oil.  . 
Model  cyl 


0. 1 

o.ois 

0.22 

0.03 

o.  I 

0.138 

0.094 

o.ts 

0.15 

0.04s 


2.42 

4.37 
2.91 
0.65 
2.17 
1-57 

2.87 

i^63 
1 .00 
1.87 


4-34 
2.42 
1.38 


0.56 
3^13 
0.90 

1^25 

4^34 
2.42 
1.87 

I^2S 


O.C75 

0.23 

0.62 

1.30 

1.56 


2.77 
0.94 


0.52 

0.88 

0.525 

0.96 

1 .  12 

0.094 

o.  14 


0.44 
0.48 


I  .  12 
0.81 


0.511 

0.69 

0.99 

0.  107 

0.639 
0.323 

0.392 

0.81 
0.229 

0.555 
0.I7I 

2  .  22 

I  .  II 

0.92 

0.65 

0.225 

0.927 

0.58 

0.835 

2.86 

1-43 

I  .III 

0.371 

1.365 

0.461 

0.075 

0.231 

1. 185 

0.927 

1 .  Ill 
0.314 
0.314 
1 .  272 
0.45S 
0.781 
0.655 
0.83s 
0.908 
0.600 
1.375- 
0.625 
0.115 
o.  169 
0.2 
0.769 
1. 15 
0.587 
1. 153 
1. 153 
0.4s 


1.79 
6.46 
4.33 
0.67 
2.23 
2.42 


2.84 
1. 14 
4.16 


15.80 
8.32 
6.90 


I 

.41 

6 

.95 

3 

.10 

6 

25 

21 

45 

10 

72 

8 

33 

2 

78 

4 

32 

0 

47 

I 

44 

8 

89 

2 

90 

3 

47 

5.86 


2  .61 

6.81 

2. 1 

6.00 

7.C4 

0.72 

1 .06 


3 

36 

4 

02 

8 

64 

6 

20 

Table  35. — Steam  Engine  Lubrication  {Power). 


ii8 


BEARINGS  AND  THEIR  LUBRICATION 


Esti- 

No. 

mated 
h.p. 

SI 

240 

52 

160 

53 

130 

54 

100 

55 

130 

56 

100 

57 

200 

58 

160 

59 

100 

60 

85 

61 

no 

62 

100 

6;j 

1^0 

64 

70 

6S 

70 

66 

100 

67 

70 

68 

70 

69 

70 

70 

70 

71 

70 

72 

50 

73 

100 

74 

50 

75 

70 

76 

70 

77 

40 

78 

50 

79 

50 

80 

85 

81 

SO 

Description  of  engine 


Cylin- 
ders 


Revolu- 

Stroke 

tions  per 

minute 

23 

106 

42 

70 

36 

90 

14 

260 

42 

75 

36 

100 

30 

80 

30 

85 

36 

90 

18 

190 

14 

212 

14 

223 

32 

85 

12 

300 

21 

i6s 

36 

80 

18 

180 

36 

90 

36 

88 

36 

8S 

34 

90 

14 

260 

20 

130 

12 

300 

30 

100 

14 

210 

12 

300 

12 

240 

IS 

170 

17 

los 

12 

250 

Steam 
pressure 


Sq.    ft. 

rubbed 

over  per 

hour 


Wright  Corliss 

Nordberg  Corliss 

Robt.  Armstrong  automatic.  .  . 

Corliss 

Corliss 

Slide  valve 

Greene 

Variable  speed,  St.  Louis  Corliss 

Atlas  automatic 

Ames 

Automatic  piston  valve 

Nordberg  Corliss 

Ideal 

Buckeye    

Corliss 

Atlas  single  valve  automatic .  .  . 

Corliss 

Bates  Corliss 

St.  Louis  Corliss 

New  Brown 

Atlas 

McEwen 

Fitchburg 

Slide  valve 

Ball 

Atlas 

Center  crank 

Slide-valve 


85 

125 


135-140 
125 


no 
100 


no  very  wet 

100 

135-140 

80 

80 

80 

90 
150 


100 
100 


140,500 

138,800 

135.800 

133.500 

132,100 

132,000 

125,800 

120,200 

118,800 

116,800 

116,600 

114.500 

113,800 

113,100 

108,900 

105,800 

101,800 

101,800 

99,500 

96,350 

96,200 

9S.400 

95.400 

94,400 

94,400 

92,500 

84.750 

75,400 

66,800 

60,900 

31.410 


Table  35, — Continued. — Steam  Engine  Cylinder  Lubrication  {Power) 


LUBRICANTS 


119 


No. 


Oil  used 


Name  or  description 


Price  per 
gal.      I 


Amount 
used  per 
hr.,  pints 


Cost  per 
hr.. 
cents 


Pints  per 

estimated 

1000  h.  p. 

hours 


Cents  per 

estimated 

1000  h.  p. 

hours 


Pints  per 
million 
sq.  ft.  rub- 
bed over 


"  Eureka  cyl. 

Capitol 

Val valine. . .  , 
"600  W."... 


"600  W." 


"600  W." 

No.  650  dark  valve  cyl 

"  600  W."  vacuum  cyl 

Capitol 

High  grade  heavy 

Dark  heavy  cyl.  with  graph. 
Capitol  cjd 


Buckeye  cyl. 


Capitol  cyl 

No.  650  dark  valve  c^l.  . 

Franklin  oil 

Capitol  cyl 

Capitol  cyl 

W.  P.  Miller's  cyl.  ccrr.p. 


Flake  graph,  with  eng.  oil. 
Porno  oil 


"600  W." 

Premium  valve  oil . 


.35 
.50 

•  75 

•  35 
•35 
■75 
•45 


O.  IDS 
O.I2S 
0.2 
0.15 
0.138 
0.06 
0.018 
0.03 
0.04s 
067 
II 
125 
115 
083 


0.049 

0.6 

0.0835 

0.2 

0.04s 

O.I 

0.055 

0.2 
0.04 

0.055 
0.083 
0.167 
0.045 
0.083 
0.125 
0.091 


0.55 

1.87 

I  .  12 
0.48 
0.45 


O.  22 
0.28 
0.50 
0.48 
0.78 


•36 


0.87 
0.28 
0.94 

o.  24 

0.87 
0.37 
0.31 


0.34 
0.94 


Averages 


0.437 

0.781 

1.538 

1-5 

1 .061 

0.6 

0.09 

0.187 

0.45 

0.788 

i.o 

1.25 

0.885 
1. 186 
1.428 
0.49 

8.57 

1. 19 
2.857 
0.643 
1.428 
1 .1 
2.0 
8 

786 
185 
175 
9 

66 
47 
82 


3.41 
14.38 
11.22 

3-71 
450 


1 .46 
2.82 
S.90 
4-37 
7.8 


5.18 


30.0 


12.48 
4.01 

13.38 
4.81 
8.74 
7.50 
4.42 


6.75 


6.19 


0.75 
0.90 
1.47 
1 .  12 
1 .04 
0.45 
o.  14 
0.2s 
0.38 
0.57 
0.94 
1.09 
1 .01 
0.73 
0.92 
0.47 
5.94 
0.82 
2.01 
0.47 
1 .04 
0.58 
2. 10 
0.42 
0.58 
0.90 
1.97 
0.60 
1.24 
2.0s 
2.90 


Table  35. — Continued. — Steam  Engine  Cylinder  Lubrication  {Power), 


I20 


BEARINGS  AND  THEIR  LUBRICATION 


Rating 

Gallons  oil  per  minute 

Pressure  per  square  inch 

Kilo-    1                !^, 

^^     |R.  p.  m.ibtep 
watts 

Guide 

Gear 

Total 

Step 

Baffler 
drop 

Pump- 
less 
line 

Gear 

and 

guides 

Unbalanced 
rev.  wt. 

Step  block 
0.  D.,  I.  D. 

750 

1500 

3750 

Sooo 

9000 

14000 

15000 

20000 

1800 
1500 
900 
750 
750 
750 
7SO 
750 

3 
4 
9 
12 
IS 
18 
20 

25 

1,5 

X.7S 

2.0 

3.0 

4.0 

4.0 

4.0 

4.0 

2.5 
2.25 
50 
6.0 
8.0 
10.  0 

X2.0 
14.0 

7 
8 
16 
21 
27 
32 
36 
43 

290 
320 
525 
825 
900 
900 
850 
950 

60 
6s 

100 
120 

.50 

ISO 
150 

.so 

350 

385 

625 

945 

1050 

1050 

1000 

1 1 00 

120 
120 
120 
120 
120 
120 
120 
120 

10800 
19700 
65000 

lOIOOO 

148000 
190000 
216000 
233000 

8x5  3/4 

11x8 

16x9   1/4 

i6xro 

i8i  X  103/4 

20x13    1/2 

21x15 

21x15 

Table  36. — ^Veetical  Steam  Turbine  Lubrication. 


FILTRATION  OF  LUBRICATING  OILS 


Oil  after  having  passed  through  machinery  bearings  seems  to  lose  a  large 
proportion  of  its  lubricating  power,  besides  being  contaminated  with  foreign 
particles,  such  as  metallic  dust,  carbon,  dirt  and  the  like.  It  is  a  difficult  matter 
to  restore  it  to  its  original  purity  and  value.  The  usual  shop  method  is  by 
filtration,  but  the  filtered  oil  is  seldom  used  alone  as  a  lubricant.  It  is  usually 
mixed  with  a  certain  amount  that  is  fresh. 

The  three  common  methods  of  filtration  are  by  capillary  action,  by  gravity 
filtering,  or  by  pressure  filtering.     Of  these  the  second  is  the  most  common. 

G.  W.  Bissell  describes  a  capillary  filter  in  the  Trans.  A .  S.  M.  E.,  Vol.  XVII, 
page  293.  "The  oil  filter  that  I  used. consists  of  two  shallow  rectangular  tin 
trays  and  some  wide  lamp-wicks.  One  of  the  trays,  slightly  smaller  than  the 
other,  is  supported  within  it  on  two  blocks,  which  raise  it  an  inch  or  so  from  the 
bottom  of  the  larger  pan.  The  wicks  are  laid  in  the  upper  pan,  so  as  to  hang 
over  its  edges  into  the  lower  pan.  The  oil  to  be  filtered  is  poured  into  the  upper 
pan.  A  drip  cock  in  the  lower  pan,  and  a  cover  for  the  whole,  complete  the 
apparatus.  No  quantitative  tests  have  been  made,  but  the  results,  as  far  as 
the  eye  can  judge,  are  good.'* 

Gravity  filtration  consists  in  passing  the  used  oil  through  mediums,  such 
as  mineral  or  animal  wool,  bone  dust,  vegetable  or  animal  charcoal,  hair  felt, 
absorbent  cotton  and  the  like.  Heat  is  usually  applied,  and  the  final  condition 
of  the  filtered  oil  depends  upon  the  cleaness  of  the  material  and  the  rate  of  its 
flow.  It  is  doubtful  if  any  of  these  materials  can  completely  remove  finely 
divided  metallic  particles. 

Pressure  filtering  is  only  used  for  very  large  quantities  of  oil. 


LUBRICANTS  121 

SIMPLE  TESTS  FOR  GREASES. 

Professor  G.  W.  Lewis  writing  in  the  American  Machinist,  gives  a  few  simple 
tests  of  greases  as  follows: 

There  are  two  general  types  of  greases;  the  so-called  mineral,  and  the 
animal  or  tallow  greases.  Mineral  greases  should  be  made  from  a  high  grade 
refined  petroleum  oil  and  so  treated  that  the  least  possible  amount  of  foreign 
substances  be  used  to  bring  it  to  a  solid  state.  Lime  soap  is  generally  used 
and  the  mixture  emulsified  by  mechanical  sitrring  or  blowing  air  through  the 
mixture.  Some  manufacturers  use  resin,  resinous  oils,  graphite,  soap-stone 
wax,  talc,  powdered  mica,  and  asbestos,  which  are  introduced  into  the  grease 
for  the  purpose  of  creating  an  artificial  body. 

Tallow  greases  should  consist  of  some  hard  animal  or  vegetable  fat,  such 
as  tallow,  with  a  small  amount  of  mineral  oil.  Here  the  mixture  is  solidified 
by  the  use  of  a  soap,  in  this  case  a  potash  soap  being  usually  used.  The  melting 
temperature  of  tallow  grease  is  from  50  to  75  degrees  lower  than  the  melting 
temperature  of  the  mineral  grease.  Some  very  inferior  greases  are  made  from 
oils  which  are  the  waste  of  soap  factories  and  sewage  disposal  systems. 

All  greases  look  good  when  the  can  is  opened  and  nearly  all  will  work  well 
on  a  bearing  for  a  short  time,  and  then  a  poor  grease  will  cause  the  bearing  to 
heat,  which  decomposes  the  grease  into  oil  and  soap,  especially  if  the  per  cent, 
of  soap  is  high,  and  impurities  have  been  used.  Any  adulterants  used  continu- 
ally accumulate  in  the  bearings  as  the  grease  melts  away,  and  cause  gumming 
and  sometimes  scoring  of  the  bearing. 

The  consumer,  to  protect  himself,  should  make  the  foUovdng  simple  tests 
on  any  grease  he  contemplates  using,  test  for  acids,  alkalies,  volatile  matter, 
filling,  water  and  gumming. 

The  test  for  free  acid  is  best  made  by  melting  a  small  quantity  of  grease  and 
applying  blue  litmus  paper.  The  test  for  alkali  is  made  in  the  same  way  by 
applying  red  litmus  paper. 

The  test  for  volatile  matter,  such  as  naphtha  or  benzine,  is  made  by  heating 
a  measured  quantity  to  150°  to  200°  F.  and  determining  by  weight  the  amount 
volatilized. 

The  test  for  filling,  or  the  amount  of  soap  or  other  material,  used  to  form  a 
body,  can  be  partially  determined  by  melting  a  test-tube  of  the  grease  to  a 
liquid  state  and  noting  the  amount  of  residue  or  cloudiness  in  the  bottom  of  the 
tube,  which  gives  an  idea  of  the  amount  of  soap  or  other  material  used  in 
bringing  the  grease  to  a  solid  state.  A  more  accurate  test  is  to  take  a  small 
sample  of  grease  and  dissolve  in  gasoline,  then  add  phosphoric  acid  in  alcohol 
solution  and  if  any  soap  is  present,  it  will  precipitate  and  give  a  relative  idea 
of  the  amount  in  the  grease 

To  determine  the  amount  of  water  in  the  grease,  and  to  see  if  it  will  "wash 


122  BEARINGS  AND  THEIR  LUBRICATION 

out, "  place  a  small  amount  in  a  linen  cloth  or  a  common  handkerchief,  fold  in 
the  form  of  a  bag,  and  place  in  water  and  knead  for  five  minutes.  In  this  way 
the  amount  of  water  present  can  be  determined,  and  also  if  the  grease  will  wash 
out,  which  is  very  important,  especially  in  paper  making  and  laundry  machines 
where  water  comes  in  contact  with  the  bearings. 

The  simplest  method  to  determine  the  amount  of  gumming  is  to  take  a 
sherry  glass,  put  in  a  few  pieces  of  copper  wire,  and  add  concentrated  nitric 
acid,  then  add  the  grease  in  finely  divided  particles  to  the  solution,  and  stir 
until  all  the  grease  has  been  broken  up.  In  the  case  of  a  tallow  grease,  the 
tallow  will  form  a  solid  crsut  on  top  of  the  liquid  in  about  an  hour;  with  a 
vegetable  oil,  the  oil  remaining  will  be  about  the  consistency  of  butter,  while 
with  a  mineral  oil,  the  oil  remains  in  a  liquid  state.  The  blackish  substance 
that  appears  below  the  oil  represents  the  impurities  that  cause  gumming  of  the 
bearing.  Next  separate  the  acid  from  the  oil  and  the  black  substance  and  add 
gasoline  to  the  latter  two.  This  washes  out  the  oil  and  separates  the  gumming 
substances  so  that  a  fair  idea  of  the  quantity  in  the  grease  can  be  obtained. 

PENNSYLVANIA  RAILROAD  COMPANY  SPECIFICATIONS  FOR  LARD  OIL 

Two  grades  of  lard  oil,  known  in  market  as  "Extra"  and  "Extra  No.  i," 
will  be  used,  the  former  principally  for  burning  and  the  latter  as  a  lubricant. 

The  material  desired  under  this  specification  is  oil  from  the  lard  of  corn  fed 
hogs,  unmixed  with  other  oils,  and  containing  the  least  possible  amount  of  free 
acid.  Also  from  October  i  to  May  i  it  should  show  a  cold  test  not  higher  than 
40°  F.  on  from  lard  of  "mast"  or  distillery  fed  hogs  does  not  give  good 
results  in  service,  and  should  never  be  sent.  Also  care  should  be  taken  to  have 
the  oil  made  from  fresh  lard.  Old  lard  gives  an  oil  that  does  not  burn  well, 
and  also  gums  badly  as  a  lubricant.  The  use  of  the  so-called  neatsfoot  stock, 
either  alone  or  as  an  admixture  in  making  the  "Extra  No.  i"  grade,  is  not 
recommended.  Neatsfoot  oil  is  used  by  the  Railroad  Company  when  the 
price  will  admit,  but  it  is  preferred  to  have  it  unmixed. 

Shipments  must  be  made  as  soon  as  possible  after  the  order  is  placed. 
Also  shipments  received  at  any  shop  after  October  i  will  be  subjected  to  cold 
test  and  rejected  if  they  fail,  unless  it  can  be  shown  that  the  shipment  has  been 
more  than  a  week  in  transit. 

Shipments  of  the  "Extra"  grade  will  not  be  accepted  which 
I.  Contain  admixture  of  any  other  oils. 
II.  Contain  more  free  acid  than  is  neutralized  by  4  c.  c.  of  alkali 

as  described  in  the  printed  method. 
III.  Show  a  cold  test  above  45°  F.  from  October  i  to  May  i. 
IV.  Show  coloration  when  tested  with  nitrate  of  silver  as  described 
below. 


LUBRICANTS  123 

Shipments  of  "Extra  No.  i"  grade  will  not  be  accepted,  which, 
I.  Contain  admixtures  of  any  other  oils. 
II.  Contain  more  free  acid  than  is  neutralized  by  20  c.  c.  of  alkali 

as  described  in  the  printed  method. 
III.  Show  a  cold  test  above  45°  F.  from  October  i,  to  May,  i. 
The  cold  test  and  the  amount  of  free  acid  must  be  determined  in  accordance 
with  P.  R.  R.  standard  methods,  printed  copies  of  which  will  be  furnished  upon 
application. 

The  nitrate  of  silver  test  is  as  follows:  Have  ready  a  solution  of  nitrate  of 
silver  in  alcohol  and  ether,  made  on  the  following  formula: 

Nitrate  of  silver i  gram. 

Alcohol 200  grams. 

Ether 40  grams. 

After  the  ingredients  are  mixed  and  dissolved,  allow  the  solution  to  stand 
in  the  sun  or  in  diffused  light  until  it  has  become  perfectly  clear;  it  is  then 
ready  for  use  and  should  be  kept  in  a  dimly  lighted  place  and  tightly  corked. 

Into  a  50  c.  c.  test-tube,  put  10  c.  c.  of  the  oil  to  be  tested  (which 
should  have  been  previously  filtered  through  washed  filter  paper),  and  5  c.  c. 
of  the  above  solution,  shake  thoroughly  and  heat  in  a  vessel  of  boiling  water 
15  min.  with  occasional  shaking.  Satisfactory  oil  shows  no  change  of  color 
under  this  test. 

UNITED  STATES  NAVY  SPECIFICATIONS 

Following  are  the  specifications  for  lubricating  oil  for  marine  machinery 
as  issued  by  the  United  States  Navy  Department,  September  13,  1906,  and 
still  in  force. 

1.  Must  be  a  properly  compounded  oil  to  form  a  homogeneous  compound 
that  will  not  separate  under  varying  temperatures,  and  must  consist  of  a  pure 
mineral  oil,  and  not  more  than  30  per  cent,  nor  less  than  20  per  cent,  of  suitable 
non-drying  fixed  oils  as  may  be  best  suited  for  lubrication. 

2.  The  compounded  oil  must  be  free  from  rosin,  tar,  drying  oils,  sulphur, 
asphaltic  or  tarry  bodies,  soaps  or  oil  thickners,  water,  grit,  dirt  or  other  sus- 
pended matter;  and  must  be  free  from  mineral  acid,  and  must  not  contain 
more  than  2  per  cent,  of  free  oleic  acid.  The  speciiEic  gravity  to  be  between 
0.915  and  0.927  at  60°  F. 

^.  As  a  lubricant  the  oil  when  tested  on  an  oil-testing  machine,  owned  and 
operated  by  the  Government,  having  a  standard  brass  bearing  of  about  9  sq.  in. 
projected  area  on  a  polished  steel  mandrel  making  about  160  revolutions  per 
minute  with  a  surface  speed  of  about  250  ft.  per  minute  must,  to  be  satisfactory, 
perform  as  follows: 


124  BEARINGS  AND  THEIR  LUBRICATION 

(a)  The  temperature  of  bearing  must  at  no  time  during  the  test  be  per- 
mitted to  exceed  130°  F.  Only  sufficient  oil  under  test  to  be  applied  to  prevent 
excessive  friction  and  heating  of  bearing. 

(b)  The  average  load  on  the  bearing  for  two  hours  must  be  at  least  300  lb. 
per  square  inch  of  projected  area  of  bearing. 

(c)  The  quotient  found  by  dividing  the  product  of  the  average  total  pressure 
on  journal  and  surface  speed  of  journal  in  feet  per  minute  by  the  weight  of  oil 
in  grains  used  for  lubrication  in  the  test  must  not  fall  below  325,000. 

(d)  The  duration  of  test  must  not  be  less  than  two  hours  nor  more  than 
two  and  one-half  hours,  and  but  one  test  will  be  permitted  for  any  one  lot  of  oil. 

4.  To  be  purchased  and  inspected  by  weight,  the  number  of  pounds  per 
gallon  to  be  determined  by  the  specific  gravity  of  the  oil  at  60°  F.  multiplied 
by  8.33  lb.,  the  weight  of  a  gallon  (231  cu.  in.)  of  distilled  water  at  the  same 
temperature. 

5.  Flashing  point  must  not  be  below  400°  F. 

6.  Freedom  from  gumming.  Using  a  half-pint  brass  oil  cup  maintained 
at  about  140°  F.,  practically  equal  quantities  of  oil  must  feed  through  the  wick 
in  equal  intervals  of  time  for  three  intervals  of  eight  hours  each;  the  wick  to  be 
of  zephyr  wool,  four  strands,  doubled  once.  The  oil  in  the  cup  will  be  brought 
to  the  original  level  at  the  beginning  of  each  hour,  and  not  less  than  14  oz. 
avoirdupois  must  feed  through  the  wick  during  the  first  period  of  eight  hours. 
At  the  end  of  test  the  wick  must  be  clean  and  the  sides  of  oil  cup  bright  and 
clean. 

7.  Cold  Test. — The  oil  must  flow  at  a  temperature  of  32°  F. 

8.  Freedom  Jrom  Acid. — A  small  quantity  when  applied  to  a  polished  copper 
plate  must  not  turn  the  surface  of  the  metal  green  if  allowed  to  stand  exposed 
to  the  air  for  twenty-four  hours. 

9.  Viscosity  by  Engler  viscosimeter.  At  90°  F.  must  not  be  below  685;  at 
150°  F.  must  not  be  below  155;  at  225°  F.  must  not  be  below  75,  compared  with 

distilled  water  (49)  at  90°  F. 

10.  Oil  accepted  and  paid  for  under  these  specifications  will  be  the  subject 
of  further  tests  in  actual  use  on  board  ship,  and  any  brand  of  oil  failing  to  afford 
thoroughly  efficient  service  in  these  conclusive  tests  shall  not  again  be  considered 
for  naval  use. 

UNITED  STATES  WAR  DEPARTMENT  SPECIFICATIONS 
No.  I 
ENGINE    OIL 

This  oil  is  suitable  for  general  lubrication  of  marine  engines.  All  work 
except  cylinder. 


LUBRICANTS  1 25 

SPECIFICATIONS 

Must  be  a  compounded  oil  composed  of  25  per  cent,  pure  acidless  tallow 
oil  and  75  per  cent  pure  mineral  oil.  Must  be  free  from  acid,  alkali,  and 
suspended  matter,  and  must  satisfactorily  pass  the  following  tests: 

TESTS 

Specific  Gravity. — Must  not  be  below  .920  nor  above  .923  at  60°  F. 
Flash. — Must  not  flash  below  415°  F. 
Fire. — Must  not  burn  below  470°  F. 
Viscosity. — Must  not  be  below  8. 11  at  50°  C.     (Engler.) 
Cold  Test. — Must  flow  at  a  temperature  of  35°  F. 

Acid. — Must  not  give  an  acid  reaction  on  polished  copper  in  24  hours. 
Alkali. — Ash  must  not  show  an  alkaline  reaction. 
Water. — Oil  must  not  froth  nor  bump  in  flash  cup  when  heated. 
Saponification  — When  oil  is  treated  with  alcoholic  caustic  potash   must 
show  presence  of  25  per  cent,  tallow  oil. 

No.  2 

CYLINDER    OIL    (dARK) 

This  oil  is  suitable  where  sieam  pressures  are  high  and  lubricating  conditions 
severe. 

SPECIFICATIONS 

Must  be  a  compounded  oil  composed  of  5  per  cent,  pure  acidless  tallow 
oil  and  95  per  cent,  pure  mineral  oil.  Must  be  free  from  acid,  alkali,  tarry  or 
suspended  matter,  and  must  satisfactorily  pass  the  following  tests. 

TESTS 

Specific  Gravity. — Must  not  be  less  than  .900  nor  more  than  .905  at  60°  F 

Flash. — Must  not  flash  below  540°  F. 

Fire. — Must  not  burn  below  600°  F. 

Viscosity. — Must  not  be'below  3.83  at  100°  C.     (Engler.) 

Acid. — Must  not  give  an  acid  reaction  on  polished  copper  in  24  hours. 

Alkali. — Ash  must  not  show  an  alkaline  reaction. 

Water. — Must  not  froth  nor  bump  when  heated  in  flash  cup. 

Tarry  and  Suspended  Matter. — Put  5  c.c.  oil  in  100  c.c.  stoppered  measuring 
cylinder  and  95  c.c.  of  benzine.  Shake  well  and  allow  to  stand  at  least  10 
minutes,  and  must  be  no  precipitation. 

Volatility. — Heat  small  quantity  of  oil  on  watch  glass  for  two  hours  at 
400°  F.     There  must  not  be  a  loss  of  more  than  5  per  cent,  by  weight. 

Saponification. — When  oil  is  treated  with  alcoholic  caustic  potash  must 
show  presence  of  5  per  cent,  tallow  oil. 


126  BEARINGS  AND  THEIR  LUBRICATION 

No.  3 

CYLINDER    OIL    (FILTERED) 

This  oil  is  suitable  where  a  light  colored  valve  oil  is  desired. 

SPECIFICATIONS 

Must  be  a  compounded  oil  composed  of  15  per  cent,  pure  acidless  tallow  oil, 
and  85  per  cent,  pure  filtered  mineral  oil.  Must  be  free  from  acid,  alkali 
tarry  or  suspended  matter,  and  must  satisfactorily  pass  the  following  tests: 

TESTS 

Specific  Gravity. — Must  not  be  less  than  .891  nor  more  than  .895  at  60°  F. 

Flash. — Must  not  flash  below  470°  F. 

Fire. — Must  not  burn  below  540°  F. 

Viscosity. — Must  not  be  less  than  2.30  at  100°  C.  (Engler.) 

Acid. — Must  not  give  an  acid  reaction  on  polished  copper  in  24  hours. 

Alkali. — Ash  must  not  show  an  alkaline  reaction. 

Water. — Must  not  froth  nor  bump  when  heated  in  flash  cup. 

Tarry  and  Suspended  Matter. — Put  5  c.c.  oil  in  100  c.c.  stoppered  measuring 
cylinder  and  95  c.c.  of  benzine.  Shake  well  and  allow  to  stand  at  least  10 
minutes.     There  must  be  no  precipitation. 

Volatility. — Heat  small  quantity  of  oil  on  watch  glass  for  2  hours  at 
400°  F.     There  must  not  be  a  loss  of  more  than  5  per  cent,  by  weight. 

Saponification. — When  oil  is  treated  with  alcoholic  caustic  potash,  must 
show  a  presence  of  15  per  cent,  tallow  oil. 

No.  4 

GAS   ENGINE    OIL 

This  oil  is  suitable  for  lubrication  of  gas  engines  that  are  water  cooled,  also 
for  bearings. 

SPECIFICATIONS 

Must  be  a  pure  filtered  mineral  oil.  Must  be  free  from  acid,  alkali,  suspended 
matter,  and  satisfactorily  pass  the  following  tests. 

TESTS 

Specific  Gravity. — Must  not  be  less  than  .896  nor  more  than  .899  at  60°  F. 

Flash. — Must  not  flash  below  400°  F. 

Fire. — Must  not  burn  below  460°  F. 

Viscosity. — Must  not  be  less  than  3.09  at  50°  C.     (Engler.) 

Cold  Test. — Must  flow  at  a  temperature  of  29°  F. 

Acid. — Must  not  give  an  acid  reaction  on  polished  copper  in  24  hours. 


LUBRICANTS  1 27 

Alkali. — Ash  must  not  show  an  alkaline  reaction. 
Water. — Must  not  froth  nor  bump  when  heated  in  flash  cup. 
Saponification. — Must  be  unaffected  by  an  alcoholic  solution  of  caustic 
potash. 

No.  5 

GAS   ENGINE    OIL 

This  oil  is  suitable  for  lubrication  of  gas  engine  cylinders  that  are  water 
cooled,  also  bearings,  where  an  oil  of  lighter  body  than  No.  4  is  desired. 

SPECIFICATIONS 

Must  be  a  pure  filtered  mineral  oil.     Must  be  free  from  acid,  alkali,  and 
suspended  matter,  and  must  satisfactorily  pass  the  following  tests. 

TESTS 

Specific  Gravity. — Must  not  be  less  than  .874  nor  more  than  .877  at  60°  F. 

Flash.— Must  not  flash  below  385°  F. 

Fire. — Must  not  burn  below  430°  F. 

Viscosity. — Must  not  be  less  than  4.61  at  50°  C.     (Engler.) 

Cold  Test. — Oil  must  flow  at  temperature  of  ^t,°  F. 

Acid. — Must  not  give  a  reaction  of  acid  on  polished  copper  in  24  hours. 

Alkali. — Ash  must  not  show  an  alkaline  reaction. 

Water. — Must  not  froth  nor  bump  when  heated  in  flash  cup. 

Saponification. — Must  be  unaffected  by  an  alcoholic  solution  of  caustic  potash . 

No.  6 

ENGINE  OIL       •      , 

This  oil  is  suitable  for  lubrication  of  high  speed  engines,  dynamos,  and 
high  speed  work  generally. 

SPECIFICATIONS 

Must  be  a  pure  filtered  mineral  oil.     Must  be  free  from  acid,  alkali,  and 
suspended  matter,  and  pass  satisfactorily  the  following  tests. 

TESTS 

Specific  Gravity. — Must  not  be  less  than  .874  nor  more  than  .878  at  60°  F. 

Flash.— Musi  not  flash  below  375°  F. 

Fire. — Must  not  burn  below  420°  F. 

Viscosity. — Must  not  be  less  than  2.40  at  50°  C.  (Engler). 

Cold  Test. — Must  flow  at  a  temperature  of  14°  F. 


128  BEARINGS  AND  THEIR  LUBRICATION 

Acid. — Must  not  give  an  acid  reaction  on  polished  copper  in  24  hours. 
Alkali. — Ash  must  not  show  an  alkaline  reaction. 
Water. — Must  not  froth  nor  bump  when  heated  in  flash  cup. 
Saponification. — Must  be  unaffected  by  an  alcoholic  solution  of  caustic 
potash. 

No.  7 

ENGINE  OIL 

This  oil  is  suitable  for  very  troublesome  bearings. 

SPECIFICATIONS 

Must  be  a  pure  filtered  mineral  oil.     Must  be  free  from  acid,  alkali,  and 
suspended  matter,  and  pass  satisfactorily  the  following  tests. 

TESTS 

Specific  Gravity. — Must  not  be  less  than  .877  nor  more  than  .880  at  60°  F. 
Flash. — Must  not  flash  below  340°  F. 
Fire. — Must  not  burn  below  390°  F. 

Viscosity. — Must  not  be  less  than  4.50  at  50°  C.  (Engler). 
Cold  Test. — Must  flow  at  a  temperature  of  32°  F. 
Acid. — Must  not  give  an  acid  reaction  on  poHshed  copper  in  24  hours. 
Alkali. — ^Ash  must  not  show  an  alkaline  reaction. 
Water. — Must  not  froth  nor  bump  when  heated  in  flash  cup. 
Saponification. — Must  be  unaffected  by  an  alcoholic  solution  of  caustic 
potash. 

No.  8 

ENGINE  OIL 

This  oil  is  suitable  for  engine  bearings  and  difficult  work  where  pressure  is 
not  extraordinarily  high. 

SPECIFICATIONS 

Must  be  a  pure  filtered  mineral  oil.     Must  be  free  from  acid,  alkali,  sus- 
pended matter,  and  pass  satisfactorily  the  following  tests. 

TESTS 

Specific  Gravity. — Must  not  be  less  than  .875  nor  more  than  .879  at  60°  F. 

Flash.— M.\isi  not  flash  below  380°  F. 

Fire. — Must  not  burn  below  420°  F. 

Viscosity. — Must  not  be  less  than  2.98  at  50°  C.  (Engler). 


LUBRICANTS  1 29 

Cold  Test. — Must  flow  at  a  temperature  of  22°  F. 

Acid. — Must  not  give  an  acid  reaction  on  polished  copper  in  24  hours. 
Alkali. — Ash  must  not  show  an  alkaline  reaction. 
Water. — Must  not  froth  nor  bump  when  heated  in  flash  cup. 
Saponification. — Must  be  unaffected  by  an  alcoholic  solution  of  caustic 
potash. 

No.  9 

MACHINE  OIL 

This  oil  is  suitable  for  shafting  and  ordinary  lubricating  duties  on  light 
running  machinery. 

SPECIFICATIONS 

Must  be  a  pure  filtered  mineral  oil.     Must  be  free  from  acid,  alkali,  and 
suspended  matter,  and  pass  satisfactorily  the  following  tests. 

TESTS 

Specific  Gravity. — Must  not  be  less  than  .868  nor  more  than  .872  at  60°  F. 
Flash. — Must  not  flash  below  370°  F. 
Fire. — Must  not  burn  below  420°  F. 
Viscosity. — Must  not  be  less  than  2.38  at  50°  G.  (Engler). 
Cold  Test. — Must  flow  at  a  temperature  of  14°  F. 
Acid. — Must  not  give  an  acid  reaction  on  polished  copper  in  24  hours. 
Alkali. — Ash  must  not  show  an  alkaline  reaction. 
Water. — Must  not  froth  nor  bump  when  heated  in  flash  cup. 
Saponification. — Must  be  unaffected  by  an  alcoholic  solution  of  caustic 
potash. 

No.  10 

MACHINE  OIL 

This  oil  is  suitable  for  ordinary  machinery  where  heavy  pressure  and  slow 
speed  call  for  an  oil  of  heavier  body  thau  No.  9. 

SPECIFICATIONS 

Must  be  a  pure  filtered  mineral  oil.     Must  be  free  from  acid,  alkali,  and 
suspended  matter,  and  pass  satisfactorily  the  following  tests. 

TESTS 

Specific  Gravity. — Must  not  be  less  than  .890  nor  more  than  .893  at  60°  F. 
Flash. — Must  not  flash  below  390°  F. 
9 


I30 


BEARINGS  AND  THEIR  LUBRICATION 


Fire. — Must  not  burn  below  440°  F. 
Viscosity. — Must  not  be  less  than  4.29  at  50°  C.  (Engler). 
Cold  Test. — Must  flow  at  a  temperature  of  32°  F. 
Acid. — Must  not  give  an  acid  reaction  on  polished  copper  in  24  hours. 
Alkali. — Ash  must  not  show  an  alkaline  reaction. 
Water. — Must  not  froth  nor  bump  when  heated  in  flash  cup. 
Saponification. — Must  be  unafi"ected  by  an  alcoholic  solution  of  caustic 
potash. 

No.  II 

ENGINE  OIL 

This  oil  is  suitable  for  use  in  cylinders  of  kerosene  engines. 

SPECIFICATIONS 

Must  be  a  compounded  oil  composed  of  20  per  cent,  pure  acidless  tallow  oil, 
and  80  per  cent,  pure  filtered  mineral  oil.  Must  be  free  from  acid,  alkali,  and 
suspended  matter,  and  pass  satisfactorily  the  following  tests. 

TESTS 

Specific  Gravity. — Must  not  be  less  than  .884  nor  more  than  .888  at  60°  F. 
Flash. — Must  not  flash  below  395°  F. 
Fire. — Must  not  burn  below  430°  F. 
Viscosity.^yL\xsi  not  be  less  than  4.06  at  50°  C.  (Engler). 
Cold  Test. — Must  flow  at  a  temperature  of  30°  F. 
Acid. — Must  not  give  an  acid  reaction  on  polished  copper  in  24  hours. 
Alkali. — Ash  must  not  show  an  alkaline  reaction. 
Water. — Must  not  froth  nor  bump  when  heated  in  flash  cup. 
Saponification. — When  oil  is  treated  with  alcoholic  caustic  potash  must  show 
presence  of  20  per  cent,  tallow  oil. 

No.  12 

CYLINDER  OIL 

This  oil  is  suitable  for  use  in  ammonia  cylinders  of  ice  and  refrigerating 
machinery. 

SPECIFICATIONS 

Must  be  a  pure  filtered  mineral  oil.  Must  be  free  from  acid,  alkali,  and 
suspended  matter,  and  satisfactorily  pass  the  following  tests. 


LUBRICANTS  131 

TESTS 

Specific  Gravity. — Must  not  be  less  than  .870  nor  more  than  .875  at  60°  F. 
Flash. — Must  not  flash  below  370°  F. 
i^fre.— Must  not  burn  below  420°  F. 
Viscosity. — Must  not  be  less  than  2.38  at  50°  C.  (Engler). 
Cold  Test. — Must  flow  at  temperature  of  10°  F. 

Acid. — Must  not  give  an  acid  reaction  on  polished  copper  in  24  hours. 
Alkali. — ^Ash  must  not  show  an  alkaline  reaction. 
Water. — Must  not  froth  nor  bump  when  heated  in  flash  cup. 
Saponification. — Must   be  unaffected  by  an  alcoholic  solution  of  caustic 
potash. 

No.  13 

CYLINDER  OIL 

This  oil  is  suitable  for  use  in  cylinders  of  Westinghouse  Engines. 

SPECIFICATIONS 

Must  be  a  pure  mineral  oil,  and  pass  satisfactorily  the  following  tests. 

TESTS 

Specific  Gravity. — Must  not  be  less  than  .892  nor  more  than  .896  at  60°  F. 

Flash. — Must  not  flash  below  450°  F. 

Fire. — ^-Must  not  burn  below  540°  F. 

Viscosity. — Must  not  be  less  than  3. 11  at  100°  C.  (Engler). 

Acid. — Must  not  give  an  acid  reaction  on  polished  copper  in  24  hours. 

Alkali  — ^Ash  must  not  show  an  alkaline  reaction. 

Water. — Must  not  froth  nor  bump  when  heated  in  flash  cup. 

Sponification. — Must  be  unaffected  by  an  alcoholic  sloution  of  caustic 
potash. 

Tarry  and  Suspended  Matter. — ^Put  5  c.c.  oil  in  100  c.c.  stoppered  measuring 
cylinder  and  95  c.c.  of  benzine.  Shake  well  and  allow  to  stand  at  least  10 
minutes;  must  give  only  slight  precipitation. 

Volatility. — Heat  small  quantity  on  watch  glass  for  two  hours  at  400°  F. 
Must  not  lose  more  than  5  per  cent,  by  weight. 


SECTION  VII 
Design  of  Sliding  Surfaces  and  Special  Bearings 

We  are  indebted  to  Professor  Sweet  for  a  most  valuable  discussion  of  the 
design  of  flat  sliding  surfaces  in  a  little  book   Things  That  Are  Usually  Wrong. 

The  following  is  adapted  from  that  source.  The  underlying  principal  is  to 
make  the  wearing  surfaces  of  equal  length.  With  this  proportion  experience 
has  shown  that  even  under  varying  loads  and  speeds  the  wear  is  very  small  and 
practically  uniform  over  the  entire  extent  of  the  contact  surfaces.  The  reason 
is  stated  thus:  ''Things  that  do  not  tend  to  wear  out  of  true  do  not  wear  much." 

Professor  Sweet  has  used  this  principle  in  the  cross-head  shoes  and  guides 
of  steam  engines,  for  the  wearing  surfaces  between  column  and  knee  and  hori- 
zontal slide  and  knee  of  a  milling  machine,  for  the  ram  and  guides  of  a  punch 
press  and  for  the  cross  slide  of  a  lathe.     In  each  case  the  design  was  successful. 


FIG.    27.      SHORT   CROSS-HEAD   GUIDE   WITH   RATCHET   SURFACES. 

Fig.  27  shows  a  cross  head  and  guide  with  equal  length  bearing  surfaces 
made  by  reducing  the  length  of  the  guide  through  cutting  a  ratchet  surface  at 
each  end.  This  surface  holds  the  oil  pushed  along  by  the  end  of  the  head  and 
acts  as  a  reservoir  from  which  lubricant  is  taken  on  the  return  stroke.  This 
construction  is  said  to  prevent  the  splashing  and  throwing  of  oil  to  a  much 
greater  extent  than  a  full  length  guide. 

In  some  classes  of  machinery,  particularly  machine  tools  where  the  sliding 
members  are  short  stroked  much  of  the  time,  excessive  local  wear  may  take 
place  either  on  the  sliding  member  or  on  the  guides.  The  shaper  is  an  illus- 
tration. Here  the  contact  surfaces  of  ram  and  guide  are  sure  to  wear  most  at 
the  front  end.  As  a  partial  remedy,  Professor  Sweet  suggests  to  cut  away  the 
wearing  surfaces  at  the  rear  increasing  the  proportion  of  gaps  to  lands  as  the 
'  132 


DESIGN  OF  FLAT  SLIDING  SURFACES 


133 


rear  end  is  approached.  Fig.  28  shows  how  this  is  done  and  gives  a  graphical 
method  of  determining  the  widths  of  gaps  and  lands.  The  shaded,  portion 
represents  the  parts  of  the  surface  that  have  been  cut  away  and  the  white  por- 
tion those  remaining  to  take  the  wear. 

The  method  of  laying  out  is  to  draw  a  diagonal  AB  across  the  surface  to  be 
relieved ;  then  lay  ofif  the  line  BC  at  a  convenient  angle  with  the  length  of  the 
guide,  say  45  degrees.     From  C  measure  CD  equal  to  the  width  of  the  guide 


C  D  O  "^ 

FIG.    28.       METHOD    OF    RELIEVING    FLAT    WEARING    SURFACES. 

and  draw  DE  parallel  to  BC  and  intersecting  AB  Sit  F.  The  vertical  distance 
from  F  to  the  upper  edge  of  the  slide  is  the  width  to  be  relieved  in  the  next 
horizontal  section  of  the  guide  having  a  length  equal  to  the  guide  width.  This 
process  of  laying  out  is  repeated  to  the  end,  the  widths  of  the  gaps  increasing. 

In  a  similar  manner  the  surface  of  the  sliding  member  should  be  relieved 
but  sloped  in  the  opposite  direction.  If  the  sliding  member  is  very  short  in 
comparison  to  the  length  of  the  guide  no  relief  will  be  necessary. 

With  regard  to  the  ways  of  lathes,  Professor  Sweet  says:  ''There  has  always 
been  diverse  practice  and,  from  time  to  time,  much  discussion  about  the  guiding 


inir     r 


^ 


s 


FIG.    29.      A   DESIGN    FOR   LATHE   WAYS. 

of  the  slide-rests  of  lathes.  The  V  has  the  merit  of  remaining  free  from  lost 
motion  however  much  worn,  but  nothing  is  more  ridiculous  than  two 
F's,  for  the  one  at  the  back  does  no  good  and  costs  money.  The  common  flat 
way  is  bad  because  the  guiding  surfaces  are  too  far  apart.  The  plan  adopted 
by  John  Lang  &  Sons,  Fig.  29,  is  much  better,  the  guiding,  though  by  flat  sur- 
faces, is  at  the  front  where  it  ought  to  be,  and  when  all  the  metal  and  work  is 
concentrated  on  the  one  guide  it  can  be  twice  as  long  and  four  times  as  efficient." 

LUBRICATION  OF  SLIDING  SURFACES 

The  lubrication  of  sliding  surfaces  is  difficult  owing  to  the  wiping  action  of 
I  he  moving  member.     It  is  common  to  oil  groove  the  surfaces  and  feed  oil  by 


134  BEARINGS  AND  THEIR  LUBRICATION 

some  of  the  ordinary  methods.  Unless  oil  is  introduced  under  pressure  the 
grooves  should  run  across  the  face  at  right  angles  to  the  direction  of  motion  and 
should  be  stopped  at  some  distance  from  the  edges.  If  thus  made  they  will  aid 
in  forming  and  maintaining  a  lubricating  film.  If  forced  lubrication  is  used 
the  grooves  can  be  cut  along  the  diagonals  of  the  surface  and  the  lubricant 
introduced  through  a  hole  at  their  intersection.  All  of  the  edges  of  the  oil 
grooves  and  the  edges  of  both  slides  and  guides  should  be  chamfered  or  rounded 
over,  to  do  away  with  scraping  action. 

Planer  ways  are  commonly  oiled  by  a  small  roller  dipping  in  an  oil  pocket, 
the  number  used  depending  upon  the  length  of  the  bed.  Professor  Sweet  says 
that  it  would  be  a  great  improvement  'Ho  use  deep  pockets  and  wheels  6  inches 
or  a  foot  in  diameter,  and  then  to  pipe  the  oil  from  the  end  pockets  to  the  bottom 
of  the  roller  pockets  and  thus  make  the  oiling  continuous." 


INFLUENCE  OF  REDUCING  THE  ARC  OF  CONTACT 

The  arc  of  contact  between  journal  and  brass  in  railroad  car  axle  bearings 
is  made  considerably  less  than  i8o  degrees.  This,  of  course,  reduces  the  pro- 
jected area  and  is  sometimes  referred  to  as  a  peculiarity  in  design.  John 
Goodman,  Proc.  Inst.  C.  E.,  Vol.  LXXXIX,  gives  the  results  of  experiments  and 
an  accompanying  empirical  mathematical  investigation  which  shows  that  the 
force  of  friction  for  a  given  set  of  conditions  can  be  reduced  by  decreasing  the 
arc  of  journal  contact  within  certain  limits. 

The  experiments  were  made  with  an  alloy  bearing  composed  of  i  part  tin 
and  7  parts  copper;  it  was  4  in.  long,  2  in.  in  diameter,  while  its  width  or  the 
length  of  the  chord  of  contact  was  varied  from  2  in.  to  1/2  in.  The  journal 
with  which  it  ran  was  made  of  manganese  steel.  Several  forms  of  lubrication 
were  tried.  The  results  show  that  the  oil  film  is  probably  at  its  best  for  an 
arc  of  contact  subtended  by  a  center  angle  of  from  80  to  no  degrees. 
Table  37  gives  the  mean  results  of  these  experiments. 
The  empirical  formula  which  agrees  very  closely  with  the  experimental 
results  is: 

Log.  R  =  log.  i?i  +0.00671  Co 
in  which  R  =  the  frictional  resistance  to  be  determined  when  the  width  of  the 

chord  is  equal  to  C. 

i?i  =  the  frictional  resistance  found  by  experiment,  when  the  width 

of  the  chord  is  C^. 

r  =  the  frictional  resistance  when  the  width  of  the  chord  =  i  (or  o.oi 

of  the  diameter).     This  is  termed  the  base  of  the  curve. 

Co=C-C,. 


REDUCING  ARC  OF  CONTACT  13$ 

To  show  the  use  of  this  equation,  two  examples  are  given  as  follows: 
Given  a  shaft  5  in.  in  diameter  fitted  with  a  semi-circular  brass,  the 
total  frictional  resistance  is  found  by  experiment  to  be  8  lb.,  it  is 
required  to  find  how  much  the  friction  will  be  reduced  by  cutting 
down  the  chord  of  the  arc  of  contact  to  3  in.,  the  total  load  is  assumed 
to  be  unaltered. 
Here  R^=S  lb.,  log.  8=0.90309. 
C    =;^  in.,  =60  units. 
0^=5  in.,  =  100  units. 
Cq  =  2  in.,  =  40  units. 
Then  log.  R  =0.90309—0.00671X40: 
=  0.63469 
i?  =4.31  lb. 
Thus,  by  reducing  the  width  of  the  brass  from  5  in.  to  3  in.,  the  frictional 
resistance  has  been  reduced  from  8  lb.  to  4.31  lb.,  or  46  per  cent. 

Given  an  axle  of  5  in.  in  diameter  carrying  a  load  of  7,500  lb.  on  a 
semi-circular  bearing  10  in.  long;  the  frictional  resistance  is  50  lb. 
Assuming  that  the  intensity  of  the  load  must  not  exceed  300  lb.  per 
square  inch  of  projected  area,  how  much  can  the  frictional  resistance 
be  reduced  by  shortening  the  arc  of  contact?  The  equation  for  the 
shortest  permissible  chord  is: 

Log.  R  minimum  =  log.  R^  +0.00671  (x—C^)  in  which  the  notation 
is  the  same  as  given  above,  except  that  x  is  the  length  of  the  shortest  per- 
missible chord. 

7.500 
Then  x  =  —  =  2.5  m.  =  50  units 

300X  10 

and  log.  R  minimum  =log.  50+0.00671  (50—  100)  =  1.3635 

R  minimum  =23.1  lb. 

Thus  the  frictional  resistance  on  this  brass  may  be  reduced  from  50  to 

23.1  lb.  without  exceeding  the  safe  intensity  of  load. 


136 


BEARINGS  AND  THEIR  LUBRICATION 


Length  of  Chord  of  Contact 

Total 

2  in. 

1.75  in. 

1.5  in. 

i.o  in. 

0.5  inch 

load 
lb. 

Angle  Subtended 

180  degrees 

122  degrees 

97  degrees 

60  degrees 

29  degrees 

P 

f 

fP 

f 

fP 

f 

fP 

f 

fP 

f 

fP 

SO 
100 
ISO 
200 
250 
300 
350 
400 
4SO 
Soo 
SSO 

0.0441 
0.0288 
0.0192 
O.OIS9 
0.0128 
0.0103 
0.0085 
0.0071 
0.0062 
0.0058 
0.0051 

2.  20 
2.88 
2.88 
3.18 
3.20 
3 -09 
2.97 
2.84 
2.79 
2.90 
2.85 

0.0458 
0.0238 
0.0160 
0.0121 
0.0097 
0 . 009 1 
0. 0070 
0.0062 
0.0055 
0.0049 
0.0045 

2.29 
2.38 
2.40 

2.42 
2.42 
2.43 
2. 45 
2.48 
2.47 
2.4s 
2.47 

0.0418 
0.0209 
0.0140 
0.0106 
0.0085 
0.0071 
0.0061 
0.0054 
0.0049 
0.0044 
0 . 0040 

2 .09 
2 .  09 

2. 10 
2. 12 
2. 12 
2.13 
2.13 
2.16 
2.  20 
2.  20 
2.20 

0.0210 
0.0105 
0.0070 
0.0055 
0 . 0044 
0.0037 
0.0033 
0.0029 
0.0028 
0.0029 
0.0027 

1.05 
1.05 
1.05 
1 .  10 
1 .  10 
I .  II 
I-I5 
1. 16 
1.26 
1.45 
1.48 

0.0126 
0.0084 
0 . 0056 
0. 0042 
0.0034 
0.0028 
0.0024 
0.0021 
0.0019 
0.0017 
0.0015 

0.63 
0.84 
0.84 
0.84 
0.84 
0.84 
0.84 
0.84 
0.84 
0.84 
0.84 

Mean 

2.89 

2.42 

2.14 

1. 18 



Conditions:     Bath  lubricator  at  104°  F..  233  r.  p.  m. 
Table  37.— Mean  Results  of  Experiments  with  Differing  Arcs  of  Contact. 

KNIFE-EDGE  BEARINGS 

Knife-edge  bearings  are  used  in  weighing  machines,  testing  machines  and 
scales.  J.  W.  Bramwell,  Engineering  News,  Vol.  LV,  page  653,  discusses 
their  design.     From  his  article  the  following  is  condensed. 

Allowable  loads  on  knife  edges  vary  with  the  manner  in  which  the  pivots  or 
knife  edges  are  held  in  the  levers  and  the  pivot  supports  of  scales,  and  secured 
to  the  bases  of  weighing  machines. 

Very  little  care  is  exercised  in  most  heavy  weighing  scalesas  to  the  proportion 
of  length  of  the  pivot  or  the  method  employed  in  securing  it  to  its  lever.  Fre- 
quently slender  pivots  are  overhung  from  the  side  of  the  lever,  and  because  of 
the  deflection,  even  under  a  moderate  load,  the  extension  of  the  pivot  beyond 
the  solid  support  is  practically  worthless.  The  case  of  the  pivot  is  much  worse 
than  the  shaft  because  the  local  distortion  is  greater;  hence  the  pivot  would 
touch  its  support  only  over  a  fraction  of  its  length. 

In  testing  machinery  greater  care  is  generally  given  to  these  details,  although 
serious  faults  are  still  prevalent  with  some  machines  considered  as  standard. 

The  testing  machine  is  nothing  but  a  weighing  scale  having  a  straining 
mechanism,  which  pulls  upon  a  specimen;  the  resistance  of  this  specimen  is 
weighed  precisely  as  upon  a  platform  scale. 


KNIFE  EDGE  BEARINGS  137 

A  defective  but  cheap  practice  in  fitting  pivots  to  levers  of  testing  machines, 
particularly  those  designed  to  carry  a  heavy  load,  is  to  cast  the  pivots  in  the 
body  of  the  lever  using  square  tool-steel  pivots  for  cores.  These  pivots  have 
a  taper  which  allows  them  to  be  driven  out  when  the  casting  is  cool.  In  spite 
of  the  care  used  in  placing  these  cores  in  the  mold  their  correct  position  will  be 
changed  after  the  casting  is  poured.  To  bring  them  into  proper  alinement  again, 
considerable  skill  and  care  is  necessary. 

The  greater  fault  of  such  practice,  however,  is  in  the  shape  of  the  hole  in  the 
casting.  The  hole  being  square  and  its  diagonal  necessarily  vertical,  a  very 
bad  breaking  point  is  made  by  the  sharp  corner.  The  tendency  to  break  is 
also  increased  by  the  chilled  iron  immediately  in  contact  with  the  pivot.  The 
initial  strain  due  to  driving  the  taper  pivot  in  still  further  reduces  the  strength 
of  the  lever  at  this  point.  The  hole  is  frequently  enlarged  by  the  wedging  action 
of  the  pivot  under  load,  and  on  account  of  the  irregular  surface  at  the  upper 
sider  of  the  pivot  it  will  assume  another  position  and  change  what  should  be  a 
fixed  ratio — the  lever  lengths.  Allowable  loads  are  variable  because  of  these 
irregularities. 

Loads  of  10,000  lb.  per  inch  of  length  on  knife  edges  are  permissible,  but 
the  pivot  must  be  flat  at  its  upper  portion,  normal  to  the  load  and  supported, 
its  whole  length  with  a  minimum  deflection  of  parts  to  ensure  reasonable 
accuracy. 

The  quality  of  steel  used  in  both  pivots  and  seats  has  an  important  bearing 
upon  the  aUowable  load.  In  all  cases  it  is  essential  that  a  high-grade  uniform 
tool  steel  be  used,  having  a  carbon  content  of  o.  90  to  i  per  cent.  Such  steel  will 
take  a  very  high  temper  and  yet  have  sufficient  ductility  to  resist  sharp  blows 
without  crumbling.  The  temper  of  the  seats  should  be  drawn  to  a  very  light 
straw  color;  that  of  the  pivots  should  be  slightly  darker. 

An  angle  of  90  degrees  for  the  knife  edge  has  given  good  results  for  heavy 
loads.  For  ordinary  weighing  machinery  and  most  testing  machinery  5,000  lb. 
per  inch  of  length  will  be  found  an  ample  allowance.  In  a  recently  built  testing 
machine  of  800,000  lb.  capacity  the  two  large  levers  had  a  load  of  10,000  pounds 
per  inch  of  length.  For  greater  loads  the  sharp  edge  is  rubbed  with  an  oilstone 
so  that  a  smoothness  is  just  visible.  A  pronounced  radius  of  knife  edge  will 
decrease  the  sensibility  of  the  apparatus.  The  seat,  if  an  angular  one,  is  shaped 
with  a  small  radius  at  the  intersection  of  the  sides  of  the  angle. 

SPINDLE  BEARINGS  FOR  TRAVERSE  SPINDLE  GRINDERS 

The  bearings  for  the  spindles  of  traverse  spindle  grinders,  essentially  watch 
factory  tools,  are  an  exception  in  that  under  proper  conditions  of  fitting  they 
are  run  without  lubrication.     Commercially  they  are  usually  run  with  a  very 


138  BEARINGS  AND  THEIR  LUBRICATION 

little  watch  oil,  benzine  or  kerosene.  They  show  a  practical  application  of  the 
phenomenon  of  air  lubrication/ 

In  diameter,  they  are  about  1/2  in.  and  run  at  speeds  from  12,000  to  13,000 
r.  p.  m.  Both  bearings  and  mating  spindles  are  of  hardened  steel  most  care- 
fully lapped  as  to  straightness  and  fit. 

W.  H.  Sawtell,  in  the  American  Machinist  for  August  5,  1909,  page  248, 
writes  of  the  conditions  under  which  these  bearings  run: 

''The  traverse  spindle  when  properly  fitted  will  run  without  a  particle  of  oil 
and  it  is  common  practice  to  use  it  in  this  way.  However,  it  must  be  kept 
clean.  Use  clean  tissue  paper  and  alcohol  to  clean  the  spindle  and  bushings, 
being  careful  to  keep  oily  fingers  from,  coming  in  contact  with  them.  If  the 
spindle  shows  a  tendency  to  cloud  up  or  stick,  rub  it  with  a  little  wad  of  paper, 
keeping  the  spindle  revolving  and  traversing  to  reach  as  much  of  the  surface  as 
possible. 

''Spindles  which  will  run  absolutely  dry  are  rare,  almost  as  rare  as  the  men 
who  understand  how  to  make  and  use  them.  Most  spindles  made  commer- 
cially are  fitted  a  little  free  and  require  the  least  bit  of  lubricating.  The  best 
method  is  to  wad  up  tightly  a  piece  of  clean  waste  about  the  size  of  a  walnut 
and  put  on  it  a  couple  of  drops  of  watch  oil.  Apply  this  to  the  revolving  spindle 
at  the  same  time  traversing  back  and  forth  several  times.  If  the  spindle  appears 
to  gum  or  cloud,  keep  at  it,  and  after  a  few  cleanings  and  oilings  it  will  settle 
down  to  business  and  give  no  further  trouble. 

"The  fit  of  these  spindles  in  their  bearings  is  very  close,  more  like  the  fit  of 
a  plug  gage  and  its  mating  ring  than  of  sliding  fits  as  we  ordinarily  know  them. 
Under  the  best  of  conditions  they  will  run  with  air  lubrication  and  after  the 
belt  is  thrown  off  will  frequently  run  on  for  two  or  three  minutes  before  stopping. 
At  the  same  time  when  standing  still  it  is  difficult  to  slide  the  spindle  through  its 
bearings  so  close  is  the  contact." 

^For  a  discussion  of  air  lubrication  see  paper  of  Professor  Kingsburry,  "Experiments  with  an 
Air-lubricated  Journal,"  published  in  the  Journal  of  the  American  Society  of  Naval  Engineers, 
Vol.  IX,  No.  2,  1897. 


SECTION  VIII 
Three  Important  Bearing  Inventions 

Before  taking  up  typical  designs  and  construction  for  journal  bearings, 
three  important  inventions  must  be  presented.  These  three  more  than  all  the 
others  have  influenced  bearing  development  not  only  in  this  country  but  also 
abroad.  The  reason  for  presenting  them  is  all  the  stronger  because  in  regard 
to  at  least  two  of  them,  there  is  a  general  misapprehension  as  to  the  real  inventor. 

Taken  in  chronological  order,  they  are  the  process  of  babbitting,  the  ball 
and  socket  seat,  and  ring  oiling. 

The  fame  of  Isaac  Babbitt  is  generally  supposed  to  rest  upon  his  discovery  of 
the  white  metal  alloy  called  to-day  babbitt  metal.  As  a  matter  of  fact  his 
invention  of  the  process  of  ''babbitting"  has  had  a  much  wider  influence  on 
the  construction  of  machinery  bearings.     While  there  are  numberless  lining 


REPRODUCTION    OF    DRAWING    OF    BALL-AND-SOCKET    BEARING    SEAT   ACCOMPANYING 
BANCROFT   OR   SELLERS   PATENT. 


metals,  many  of  which  are  far  different  from  his  original  formula,  the  majority 
of  machinery  bearings  made  to-day  consist  of  a  soft  metal  lining  anchored  by 
recesses  into  a  shell. 

In  a  little  booklet  issued  by  the  inventor  in  1848  he  says:  "Isaac  Babbitt 
would  inform  the  public  that  his  patent  does  not  consist  in  the  use  of  soft  metal, 
simply,  but  in  the  mode  of  Us  application  and  confinement  in  boxes  prepared  for 
the  purpose." 

Babbitt's  patent  was  No.  1252,  issued  July  13,  1839.  I^  was  apparently 
very  profitable.  The  Baltimore  &  Susquehanna  Railroad  Company  paid  him 
$1,050  for  the  right  to  use  it  on  their  machinery,  including  locomotives  and  cars. 

139 


I40  BEARINGS  AND  THEIR  LUBRICATION 

On  September  20,  1842,  the  secretary  of  the  United  States  Navy,  acting  under 
authority  of  Congress,  paid  him  $20,000  for  a  license  to  use  the  patent  in  the 
construction  of  bearings  for  the  machinery  of  war  vessels. 

The  second  of  these  great  inventions  is  the  so-called  Sellers'  bearing,  in- 
vented by  Edward  Bancroft.  Fig.  30  is  reproduced  from  the  patent.  The  fol- 
lowing is  taken  from  a  letter  written  by  Coleman  Sellers,  Jr. : 

"The  Bancroft  patent  for  this  type  of  hanger  was  originally  granted  October  9,  184Q. 
After  the  death  of  Mr.  Bancroft,  it  v/as  surrendered  and  re-issued  May  12,  1857,  the  re-issue 
number  is  463. 

"Mr.  Bancroft  was  in  partnership  with  the  late  William  Sellers  under  the  firm  name  of 
Bancroft  &  Sellers.  They  were  manufacturers  of  shafting  and  mill  gearing,  and  also  makers 
of  lathes,  planers  and  other  machine  tools.  This  patent  covers  the  idea  of  forming  a  portion 
of  each  box  of  spherical  form  and  supporting  them  in  guiding  sockets  so  'as  to  allow  the  boxes 
thus  held  to  play  within  the  sockets  in  the  manner  of  a  universal  joint  substantially  as  described.' 

"Mr.  Bancroft  took  out  another  patent  May  22, 1849,  '^^  which  the  box  was  supported  on 
trunion  screws  in  a  vibrating  yoke." 

The  third  of  there  inventions  is  that  of  ring  oiling.  The  inventor  is  John 
E.  Sweet,  and  the  device  was  never  patented,  although  it  is  universally  used 
to-day  in  the  bearings  of  high-grade  machinery,  particularly  electric  motors 
and  generators. 

Professor  Sweet  writes  as  follows  in  regard  to  its  early  development: 

"The  oil  ring  was  first  shown  and  described  in  Engineering  (London)  for  January,  1868, 
page  44.  The  description  is  this:  'A  self-lubricating  journal  box  is  shown  in  which  it  is  pro- 
posed to  return  the  oil  from  the  drip  cup  to  the  shaft  by  means  of  two  loose  rings.' 

"About  all  that  I  can  say  is — that  when  I  returned  from  England,  I  went  into  the  machine 
shop  at  the  Navy  Yard  and  saw  on  the  floor  a  half  dozen  or  more  hangers  made  from  the 
drawing  in  Engineering,  and  I  learned  that  they  were  to  be  sent  to  some  other  Navy  Yard, 
possibly  it  was  Beaufort. 

"When  I  reached  home  I  found  my  brother  was  putting  up  a  lineshaft  and  also  using  the 
same  thing.  When  I  built  the  first  Straight  Line  engine  and  all  afterward,  it  was  one  of  the 
features,  but  so  far  as  I  know,  no  one  else  ever  used  it  until  we  exhibited  an  engine  at  the 
Mechanics'  Fair  in  New  York,  when  Mr.  Edison  got  on  to  it  and  began  using  them  in  his 
dynamos;  after  that  it  began  to  be  used  by  others  in  various  places. 

"Had  I  taken  out  a  patent  at  first  it  would  have  expired  before  anyone  else  adopted  it,  and 
like  many  another  patent  it  would  not  have  been  worth  the  paper  it  was  written  upon." 

It  is  not  uncommon  to  find  bearings  using  all  three  of  these  inventions. 


SECTION  IX 

Typical  Designs  and  Constructions 

This  section  presents,  by  line  drawings  and  a  few  halftone  engravings,  a 
number  of  typical  bearing  designs,  constructions  and  lubricating  systems  that 
are  representative  of  American  practice.  In  the  space  available,  it  is  obviously 
impossible  to  present  more  than  a  very  few  such  illustrations. 


ELECTRIC  MOTOR  AND  GENERATOR  BEARINGS 

Figs.  31  to  37,  inclusive,  show  several  types  of  electrical  generator  and 
motor  bearings.  Figs.  31,  32,  and  33  are  from  the  General  Electric  Company. 
The  first  shows  a  large  spherical-seated,  babbitt  lined,  oil-ring  lubricated  bear- 


.-->, 


Momiual 

Dimensions 

A 

B 

c 

D 

E 

F 

// 

J 

K 

L 

A^ 

P 

R 

S 

7 

21 

H 

22 

U^ 

i'/i 

I 

VA 

12^4 

\Q% 

•lii 

5% 

6H 

2}i 

8 

24 

H 

25 

16>i 

e^4 

Wa 

i% 

14 

12  >, 

3 

(iH 

1H 

2% 

0 

27 

H 

28 

18}^ 

7H 

i% 

SVa 

15^ 

14 

M 

7 

8% 

3 

FIG.   31.      DESIGN   OF   3:1    SPLIT   SPHERICAL   SEAT   BEARINGS. 


142 


BEARINGS  AND  THEIR  LUBRICATION 


ing,  for  nominal  diameters  of  from  7  to  9  in.  This  type  is  also  used  for  much 
smaller  bearings.  Fig.  32  shows  a  pedestal  box,  while  Fig.  ;^t,  shows  a  form 
of  3  to  I  bronze  sleeve  bearing  for  small  journal  diameters. 

Figs.  34  to  37,  inclusive,  are  drawn  from  the  practice  of  the  Westinghouse 
Electric  and  Manufacturing  Company,  and  need  no  detailed  description.     A 


size  of 
Bearing 


22H 


27  H 


2m 


3>J^. 


4»/l6 


4% 


29%  5% 


H 


2h 


M 


N 


Dowel 
Pins 


Foot 
Bolts 


m-5 


FIG.    32.      DESIGN  FOR  3:1    SPHERICAL   SEAT  PEDESTAL   BEARINGS. 

feature  worth  considering  in  Fig.  37  is  the  cover  of  the  bearing  housing  permit- 
ting inspection  of  the  oil  ring  and  the  condition  of  the  lubrication.  This  cover 
is  under  spring  control,  is  locked  by  a  winged  nut  and  provided  with  packing 
to  make  a  dust-proof  joint.  The  oil  reservoirs  of  all  these  bearings  have 
drainage  plugs. 


HEAVY  GRINDER  SPINDLE  BEARINGS 

Fig.  38  shows  a  typical  construction  for  heavy  grinding  spindle  bearings 
from  the  Norton  Grinding  Company.     The  wipers  are  of  felt.     In  more  recent 


TYPICAL  DESIGNS  AND  CONSTRUCTIONS 


143 


constructions  wood  has  been  substituted,  as  it  is  believed  that  fibers  of  the  felt 
tend  to  be  picked  up  and  carried  into  the  bearing. 

FLOODED  LUBRICATION  IN  MACHINE  TOOL  PRACTICE 

The  recent  designs  of  vertical  turret  lathes  built  by  the  Bullard  Machine 
Tool  Company  have  a  system  of  "continuous  flow"  or  flooded  lubrication. 
In  the  base  of  the  machine  is  an  oil  reservoir  to  which  all  of  the  oil  drains  after 
passing  through  the  various  bearings  and  gear  boxes. 

t)  r         »i'jf<-Uril 


Nominal  Dia. 
of  Bearing 

6 

Dimensions 

Dimensious  1 
of  Set  Screwl 

A 

B 

c 

D 

E 

F 

G 

H 

J 

K 

L 

M 

N 

0 

P 

R 

S 

IH 

IH 

A 

SH 

6 

% 

2H 

% 

IH 

''/n 

y* 

'Ke 

IK, 

H 

H 

J«-ll 

H 

% 

2 

A 

a 

6?4 

% 

3 

%: 

\% 

'h2 

'A, 

'■v.. 

i\ 

Vfi 

Ji 

^/^-ll 

% 

% 

2}i 

A 

m 

-i^i 

% 

m 

%; 

\^-i 

'h, 

K« 

'A 

1% 

%. 

>i 

5i-ll 

H 

% 

2ii 

A 

l\i 

8^i 

% 

3h 

% 

' 

V4 

^'/32 

M 

1 

2h, 

^16 

^a 

94-10 

% 

« 

2h 

A 

Wi 

9 

% 

3% 

% 

2 

•Vsi 

^ 

1M« 

2Ji 

^6 

'•Jf6 

9i-10 

% 

H 

3 

B 

9 

10 

^ 

4V4 

% 

m 

H 

'Vn 

% 

\H 

2^-lc 

?ra 

7i 

?i-io 

% 

H 

■m 

B 

iQ% 

U^ 

H 

4H 

^I'a 

VA 

H 

''M 

% 

Wi 

2% 

^6 

% 

?4-10 

% 

« 

4 

B 

12 

13 

H 

5'-^ 

H 

2H 

I 

'V^ 

'ke 

m 

^H 

y* 

7^ 

?i-io 

H 

« 

4H 

B 

13M 

UVz 

H 

6^4 

H 

2% 

1 

'■y-n 

H 

V/i 

3^ 

H 

1 

1-8 

''^i'« 

M 

FIG.    33.      DESIGN    FOR  3:1    BRONZE   SLEEVE   BEARINGS. 


Submerged  therein  and  driven  from  the  driving  shaft  is  a  geared  pump 
which  delivers  the  oil  to  a  distributing  reservoir  located  on  the  outside  of  the 
column  at  a  height  to  give  a  slight  head  for  the  free  flow  of  oil  through  pipes 
leading  to  the  bearings  and  gear  chambers.  The  excess  oil  pumped  to  the 
distributing  reservoirs  drains  to  the  sump  through  an  overflow. 


144 


BEARINGS  AND  THEIR  LUBRICATION 


In  the  pipes  leading  out  of  the  reservoir  are  sights  so  that  the  flow  can  be 
observed,  and  if  a  stoppage  occurs  the  overflow  from  this  point  at  once  calls 
attention  to  it. 


FIG.    34.      MOTOR    RING   OILING   BEARING. 


FIG.    35.      RING-OILING    MOTOR   BEARING   SHOWING    OIL   RINGS   AND   GROOVES. 


FIG.    36.      RING-OILING    MOTOR  BEARING   SHOWING   OIL   RINGS  AND    GROOVES. 

The  flow  of  oil  is  at  the  rate  of  0.08  gal.  per  square  inch  of  projected  bearing 
area  per  minute. 


TYPICAL  DESIGNS  AND  CONSTRUCTIONS 


145 


Fig.  39  shows  the  table  spindle  bearing  and  method  of  lubricating  of  one 
of  these  machines. 

MACHINE  TOOL  BEARINGS 

The  illustrations  from  machine  tool  practice,  Figs.  40  to  48,  inclusive,  are 
from  the  Pratt  and  Whitney  Company  and  show  bearings  from  a  number  of 
different  machines. 


FIG.    37.      RING-OILING   MOTOR   BEARING — RIGID   SLEEVE   TYPE, 


FIG.    38.       BEARING    CONSTRUCTION    FOR   HEAVY    GRINDER   SPINDLE. 

Fig  40  is  a  section  of  the  front  spindle  bearing  of  a  turn-table  lathe.     It  is 
of  bronze,  tapered,  split,  with  an  adjusting  nut  on  each  end,  an  oil  hole  top  and 
10 


146 


BEARINGS  AND  THEIR  LUBRICATION 


bottom  with  a  longitudinal  groove  running  through  each.  These  holes  connect 
with  an  annular  chamber  in  the  headstock  casting. 

Practice  limits  the  rubbing  velocity  for  a  bearing  of  this  kind  to  800  ft.  per 
minute. 

Fig.  41  shows  a  spindle  bearing  of  a  2  1/2X26  in.  turret  lathe.  It  is 
babbitted  with  a  method  of  anchoring  the  babbitt  which  is  used  generally  on 


FIG.   39.      TABLE   SPINDLE   BEARING   OF   VERTICAL   TURRET   LATHE. 

Pratt  and  Whitney  tools.  This  consists  in  turning  a  dovetailed  groove  at  each 
end  of  the  bearing  housing  and  then  drilling  a  number  of  holes  starting  in  the 
surfaces  where  the  bearing  is  split  in  such  a  manner  that  there  holes  break 
through  into  the  lining  space  at  the   joining  surface  of  the  cap  and  bottom 


TYPICAL  DESIGNS  AND  CONSTRUCTIONS 


147 


FIG.    40.       FRONT    SPINDLE    BEARING    OF 
TURN-TABLE    LATHE. 


FIG.    41.       FRONT    SPINDLE    BEARING    OF 
TURRET   LATHE. 


f^^^^^m^ 


FIG.    42.      DETAIL   OF  ANCHORAGE   FOR   BABBITT   LINING. 


FIG.  43.      FRONT  SPINDLE  BEARING  OF  FIG.    44.      FRONT   SPINDLE   BEARING   OF 

14-INCH   TOOL-MAKER'S    LATHE.  HAND    MILLER. 


148 


BEARINGS  AND  THEIR  LUBRICATION 


half  of  the  box.  The  babbitt  enters  these  holes  and  is  thereby  held  securely 
at  the  edges  of  the  split,  at  the  same  time  due  to  its  contraction,  it  is  drawn 
tightly  into  the  dovetailed  grooves  at  the  ends  of  the  box,  see  Fig.  42. 

The  front  spindle  bearing  of  a  14  in.  tool-maker's  lathe  is  shown  in  section 
in  Fig.  43.  It  is  babbitted  with  the  regular  method  of  anchorage.  The 
proportions  are  a  little  less  than  1:2,  the  dimensions  being  2  7/16X4  in. 


Babbitt 
Thrust  Collar. 


Babbitt  Lining:. 


Ball 
Thrust  Bearing. 


FIG.    45.      SPINDLE   AND    BEARING    OF    VERTICAL    SURFACE    GRINDER. 

Fig.  44  shows  the  construction  of  the  front  spindle  bearing  of  a  No.  10 
hand  miller.  This  is  of  bronze  with  the  same  device  for  taking  up  wear  as  in 
Fig.  40.     The  bearings  are  also  made  babbitted. 

The  spindle  and  bearings  of  a  vertical  surface  grinder  are  shown  in  Fig.  45. 
Here  the  lower  bearing  is  babbitt  lined  with  ball  thrust  collars,  and  the  upper 
is  also  babbitt  lined  with  a  babbitt  thrust  collar.     Instead  of  anchoring  by 


TYPICAL  DESIGNS  AND  CONSTRUCTIONS 


149 


drilling  a  number  of  small  holes  at  the  split  surfaces,  a  number  of  l^rge  holes 
are  drilled  through  the  cap  and  its  mating  half.  These  serve  as  anchorage 
with  the  dovetailed  grooves  at  the  ends. 

The  front  spindle  bearing  of  a  5X48  in.  cylindrical  grinder  is  shown  in 


V\"\\V\^V>J  '  ^ 


FIG.  46.      FRONT  SPINDLE  BEARING  OF  5  X  4S-INCH  FIG.    47.       FRONT   SPINDLE    BEARING 

CYLINDERICAL   GRINDER.  '  OF   SPLINE   MILLER. 

section  in  Fig.  46.  It  is  a  bronze  sleeve  with  oil  grooves  top  and  bottom.  In 
the  driving  shaft  bearings  of  this  machine  felt  plugs  are  inserted  in  the  bottom 
of  the  sleeve  and  the  Osgood  system  of  grooving  is  used. 

The  front  spindle  bearing  of  a  spline  miller  is  shown  in  Fig.  47.     It  is  of 


FIG.    48.       SPINDLE    BEARING    OF   CUTTING-OFF    MACHINE. 

hardened  and  ground  tool  steel.  The  rear  bearing  is  bronze.  The  spindle  is 
likewise  of  hardened  and  ground  tool  steel  with  a  clearance  of  0.0005  in.  on  the 
diameter.     The  housing  has  a  chamber  for  oil. 

A  large  cast-iron  spindle  bearing  of  a  cutting-off  machine  is  shown  in 


I50 


BEARINGS  AND  THEIR  LUBRICATION 


Fig.  48,  and  is  of  interest  as  being  an  example  of  a  cast-iron  spindle  running 
in  a  cast-iron  box.     The  speed  is  slow. 

BEARINGS  FOR  STEAM  TURBINES 

The  step  bearings  of  vertical  Curtis  turbines  are  the  most  prominent  ex- 
ample of  the  use  of  a  cast-iron  step  for  heavy  pressures  and  high  speeds.  Table 
36  page  120  gives  the  range  of  pressures,  speeds,  and  quantities  of  oil  used. 
The  bearing  plate,  or  lower  block,  is  of  cast-iron  rigidly  held  by  the  frame. 
See  Fig.  49.  The  block  is  guided  at  the  sides  and^  carried  on  a  large  screw, 
passing  through  a  steel  nut  and  coming  in  contact  with  a  steel  block  set  in  the 


FIG.  49. 


Oil  or  Water 
STEP   BEARING    FOR   VERTICAL   CURTIS   TURBINE. 


bearing  plate.  It  is  essential  that  this  plate  should  be  rigidly  held,  that  its 
upper  face  should  be  a  true  plane  set  at  right  angle  to  the  shaft  axis,  and  that 
the  clearance  should  be  so  small  that  the  relative  alinement  of  blocks  and 
shaft  cannot  vary  appreciably. 

The  step  plate  is  likewise  of  cast-iron  and  keyed  to  the  lower  end  of  the 
shaft.  Both  plates  are  recessed  so  that  the  surfaces  of  contact  are  collars. 
Directly  above  the  step  plate  is  a  cylindrical  guide  bearing.  The  contact  faces 
of  the  blocks  must  be  truly  parallel,  and  the  contact  surfaces  of  the  end  of  the 


TYPICAL  DESIGNS  AND  CONSTRUCTIONS 


151 


screw  beneath  the  lower  block  and  its  mate  must  be  likewise  true  and  free  from 
convexity. 

Oil  is  introduced  through  the  center  of  the  screw,  passes  upward,  enters 
the  recess  in  the  center  of  the  plate,  passes  out  between  the  contact  surfaces, 
and  ascends  upward  through  the  guide  bearing,  as  indicated  by  the  arrows  in 
the  illustration. 

These  bearings  can  be  run  with  either  oil  or  water  as  the  lubricant.  In 
service  the  plates  are  actually  separated  by  a  lubricating  film;  some  four  or  five 
times  as  much  lubricant  as  is  necessary  is  usually  pumped  as  a  safeguard. 
The  greater  the  quantity  the  greater  the  separation  between  the  plates  and  the 


Upper 
Half 

Section  A-A 


iBearing  Developed  shovv- 
Oing  Oil  Grooves  and  Cooling  Coils. 

FIG.    50.      WATER-COOLED    BEARING    FOR  HORIZONTAL   TURBINE. 

less  the  danger  of  cutting  out.  The  actual  separation  of  the  plate  is,  of  course, 
only  a  few  thousandths  of  an  inch.  From  this  fact  it  can  be  seen  that  this  type 
of  bearing  calls  for  the  best  of  workmanship. 

Fig.  50  shows  a  water-cooled  bearing  for  Curtiss  horizontal  turbines 
including  the  arrangement  of  the  cooling  pipes  and  the  oil  grooving.  The 
bearing  is  provided  with  a  coil  of  thin  copper  tubing  cast  into  the  babbitt 
lining  as  close  to  the  contact  surface  as  possible,  and  having  at  the  ends  steel 
blocks  securely  brazed  on.  Water  is  led  in  through  pipes  passing  through 
stuffing  boxes  in  the  bearing  standard.  In  this  way  the  heat  of  the  bearing 
is  taken  up  by  the  cooling  water  at  the  point  where  it  is  generated.  There 
are  no  cooling  coils  in  the  oil  tanks. 

The  top  half,  or  cap  of  the  bearing,  is  cut  away,  except  at  the  ends  and 
center.  This  relieving  is  about  0.020  in.,  so  that  the  oil  fed  to  the  bearing 
fills  the  space  between  the  shaft  and  lining. 

The  oil  grooving  in  the  lower  half  of  the  bearing  is  one  straight  groove  into 
which  oil  is  fed  through  a  central  hole  as  indicated  in  the  cut,  or,  from  several 


152 


BEARINGS  AND  THEIR  LUBRICATION 


holes  leading  to  a  header  hole,  which  in  turn  is  fed  from  the  pipe  in  the  standard. 
This  groove  is  well  rounded  over  on  the  side  toward  which  the  shaft  rotates; 
on  the  side  toward  the  bottom  of  the  bearing  it  is  gradually  tapered  up  to  the 
bearing  surface.  This  assists  the  action  of  the  shaft  in  pulling  oil  into  the  region 
of  greatest  pressure. 

The  ends  of  the  main  oil  groove  terminate  in  very  small  grooves  which  run 
through  to  the  ends  of  the  bearing.  This  allows  for  the  washing  out  of  small 
particles  of  grit  and  dirt,  instead  of  letting  them  remain  at  the  ends  of  the  bear- 
ing and  cut  the  shaft  at  that  point. 

The  oil  is  discharged  from  both  ends  around  the  entire  circumference  and 
particularly  at  the  point  where  the  ends  and  middle  bands  on  the  upper  half  of 
the  bearing  are  cut  away. 


FIG.    51.       FLOATING    SLEEVE    TYPE    OF   HORIZONTAL   TURBINE    BEARING. 

At  the  ends  of  the  bearing  the  casting  is  allowed  to  project  slightly  and  a 
sheet  brass  plate  is  attached  running  down  to  a  small  clearance  around  the  shaft. 
In  this  recess  so  formed  a  deflector  or  fan  is  provided,  which  throws  off  the  oil 
from  the  ends  of  the  bearings,  preventing  it  from  leaking  out  at  the  ends  of  the 
standard  and  from  being  slooped  over  the  inside  of  the  standard  and  fmally 
seeping  out  of  the  joints. 

Bearings  of  this  type  are  designed  to  run  with  an  oil  pressure  of  from  15  to 
25  per  square  inch.  The  water  supply  pipes  are  provided  with  stop  cocks  to 
regulate  the  amount  delivered  and  the  discharge  is  open.  Water  under  low 
pressure  as  generally  supplied  from  city  mains  is  used.  Most  bearings  of  this 
kind  are  arranged  so  that  the  temperature  of  the  oil  can  be  taken  and  the 
amount  of  circulating  water  varied  to  meet  the  conditions  of  running. 

Fig.  51  shows  a  type  of  floating  sleeve  bearing  used  on  smaller  Westinghouse- 
Parsons  turbines  running  at  3600  revolutions  per  minute.     The  ratio  of  diam- 


TYPICAL  DESIGNS  AND  CONSTRUCTIONS 


153 


eter  to  length  runs  from  1:2  1/2  to  1:3.  The  bearing  proper  consists  of  a 
bronze  sleeve  provided  with  an  oil  groove  at  the. top  and  prevented  from  turn- 
ing by  a  tail  projecting  upward  into  a  recess  in  the  cap.  Around  this  tube  are 
three  concentric  sleeves,  each  diametral  clearance  being  from  0.004  to  0.006  in. 
Each  has  four  axial  rows  of  holes  to  distribute  oil  and  thus  maintain  a  film 
between  each  pair  of  mating  surfaces.  This  construction  cushions  the  shaft 
and  allows  it  to  rotate  around  its  gravity  axis  rather  than  its  geometric  axis  if 
the  two  are  not  coincident. 

The  sleeves  are  held  in  place  by  a  nut  which  in  turn  is  checked  by  a  small 
set  screw.     The  tube  with  its  nest  of  sleeves  is  held  in  a  cast-iron  shell  having 


m 


^(F~~3)^ 


FIG.    52. 


BABBITT   LINED   HORIZONTAL   TURBINE    BEARING   ARRANGED   FOR 
FLOODED    LUBRICATION. 


steel  blocks  resting  in  a  cast-iron  spherical  seating  ring  in  the  pedestal.  Under 
the  blocks  are  thin  liners  used  to  aline  the  bearing.  The  adjustment  is  in 
increments  of  0.005  ^^ch. 

Oil  under  a  slight  pump  pressure,  2  to  5  lb.,  enters  the  bearing  at  the  end, 
passes  through  it,  and  drains  into  a  well  in  the  pedestal. 

In  some  cases  the  bearing  tube  has  been  lined  with  babbitt. 

In  the  larger  machines  running  at  1800  revolutions  per  minute  or  less  a 
cast-iron  shell,  babbitt  lined  bearing  is  used  without  any  sleeves.  The  ratio 
of  diameter  to  length  is  from  1:2  to  1:2  1/2.     See  Fig.  52. 


154 


BEARINGS  AND  THEIR  LUBRICATION 


It  is  split  longitudinally  in  the  usual  manner  and  is  provided  with  spherical 
seating  blocks  as  described  above. 


3'i  Stop 
Check  Valve- 


m  Stop 
Valve. 


3H  Stop 
Valve. 


Relief  Dis- 
charge to 
Suction. 


Air  Chamber 


Strainer' 


FIG     53.      TYPE  PLAN  OF  FORCED  LUBRICATION  FOR  RECIPROCATING  ENGINES  OF  UNITED  STATES 
BATTLESHIPS   AND   ARMORED    CRUISERS. 


At  the  bottom  oil  enters  a  copper  tube  that  is  set  in  a  recess  in  the  shell  and 
held  by  the  lining;  it  divides  into  two  streams,  passes  around  the  bearing  and 


TYPICAL  DESIGNS  AND  CONSTRUCTIONS 


155 


into  a  longitudinal  groove  in  the  top  of  the  cap  cut  to  within  about  3/8  in.  of 
the  ends.  It  escapes  at  the  ends  of  the  bearing  and  drains  to  a  well  in 
the  pedestal.  The  babbitt  is  bored  out  so  that  the  horizontal  dimension  of 
the  bore  is  greater  than  the  vertical  and  thus  greater  than  the  shaft  diameter. 
The  fit  is  over  about  two- thirds  of  the  bottom  half  of  the  circumference.     This 


FIG.    54.      BALL-AND-SOCKET   RING-OILING   LINE   SHAFT   BEARING. 


FIG.    55.      STEP  AND   GUIDE   BEARINGS   FOR   VERTICAL   SHAFT. 

prevents  all  side  binding  on  the  journal,  does  away  with  any  edges  to  scrape  off 
the  oil  and  provides  a  small  reservoir  of  oil  within  the  bearing. 

Flooded  lubrication  is  used  at  low  pressures  as  in  the  other  type.  The 
oil  is  cooled  by  passing  it  through  pipes  submerged  in  water  or  through  a  tank 
containing  cooling  coils. 

At  the  open  ends  of  the  bearing  housings  are  baffles  that  prevent  leakage. 


156 


BEARINGS  AND  THEIR  LUBRICATION 


The  clearance  between  the  shaft  and  the  knife  edges  of  these  thin  collars  is 
from  0.020  to  0.030  in. 

Albert  E.  Guy  gives  both  the  European  and  standard  practice  in  regard  to 
the  design  of  De  Laval  steam  turbine  bearings  in  Tables  38  and  39.  This 
is  of  interset  and  value  because  of  the  extremely  high  speeds.  The 
bearings  are  made  of  brass  shells  lined  with  babbitt.  The  pinion  shaft 
bearings  are  lubricated  by  sight  feed,  or  wick  feed.  The  power  shaft  bearings 
are  ring  oiled  with  one  ring  in  the  middle  of  each  bearing. 


H.  r. 

R.P.M. 

d 

h 

h 
d 

Pi 

Ai 

I2 

I2 

P2 

A2 

V 

30 

20700 

0.788" 

4.56- 

5.75 

17.8 

1.73 

3  .  94" 

s. 

13.36 

1.3 

ft.  per  sec. 
71. 1 

55 

16600 

I . 102" 

5.9." 

5.36 

19.25 

2.  10 

5.67" 

5. 15 

13.36 

1.45 

79.7 

7  5 

16700 

1.181" 

8.27" 

7. 

17.64 

2.05 

7.29" 

6.17 

13.36 

..S6 

85.9 

110 

13000 

1.575" 

9.05" 

S.75 

.7.8 

2.17 

8. 11" 

5.15 

13.36 

1.62 

89.2 

150 

13000 

1.575" 

11.42" 

7.25 

19.2 

..34 

10.24" 

6.5 

14.22 

i.8s 

89.2 

225 

1 1000 

1.968" 

12.21" 

6.2 

19.2 

2.47 

10.83" 

5.5 

14.22 

1.85 

,4.S 

300 

ro6oo 

2.16s" 

14-57" 

6.72 

19.2 

2.63 

13- 11" 

6.0s 

14.  22 

1.94 

100.4 

450 

10600 

2.16s 

21.66 

10. 

19.3s 

2.64 

17.91" 

8.27 

15.65 

2.13 

100.4 

d  =  diameter  of  journal  in  inches. 

1  =  length  of  journal  in  inches. 

pi  and  P2  =  pressure  per  square  inch  of  projection  dl. 

A I  and  A2  =  coefficients 

V=  peripheral  velocity  of  journal  in  feet  per  second.     When  V  =  qo  ft.  per  second  forced 

lubrication  must  be  used— [oil  brought  to  bearing  under  pressure]. 

P I  and  P2  =  total  pressure  on  bearing  in  lb.  =  p  X  d  X 1. 

Load  on  middle  bearing  assumed  to  be  i  1/2  times  that  on  either  of  the  others. 

PX  R.P.M 

A.  = -.   A  should  ordinarily  not  exceed  2. 

168,000X1 


Table  38— European  Bearing  Practice  with  DeLaval  Single  Gear  Turbine:  One; 
Pinion  Shaft,  One  Power  Shaft. 


TYPICAL  DESIGNS  AND  CONSTRUCTIONS 


157 


Pinion  Shaft. 


Power  Shaft. 


T 

.   t2    ' 


H.  P. 

d.      ., 

I2 
d 

R.P.M. 

Peripheral 
velocity 

ds 

I3 

d3 

R.P.M. 

Peripheral 
velocity 

7 

0.395" 

31" 

^!? 

s.s.-^ 

7.2 

3000c 

ft.  per  sec. 
SI. 7 

0 
§ 

ll 

a 
0 

i" 

3tV 

3.06 

3000 

ft.  per  sec 
13. 1 

10 

0.395" 

33" 

^H" 

8.55 

7-2 

24000 

41.35 

U' 

3tV 

3.06 

2400 

10.47s 

IS 

0.532" 

4|" 

sU" 

7.75 

6.8s 

2400c 

55,7 

1.3" 

3if 

2.936 

2400 

13.62 

20 

0.708" 

411" 

4f" 

6.62 

6 

2000c 

61.8 

1-31" 

4r 

2.435 

20G0 

11.43 

30   !  0.708"  ,  si" 

5" 

7.77 

7.06 

2000c 

61.8 

1.46" 

StV 

3.S55 

2000 

12.75 

S"> 

0.871" 

6i" 

6^ 

7.89 

7.89 

20000 

76 

2" 

6i" 

4.2s 

iSoo        13.1 

75 

t" 

5i" 

4i" 

5.5 

4.875 

16400 

71.5 

11 
1 

c 

0 

1  If" 

6" 

3.43 

1500         11.45 

1 10 

1.25" 

7r 

St" 

6 

4-3 

13000 

70.9 

Nr 

6" 

2-4 

1200  1        13. 1 

150 

1.25"     j      7V 

sr 

6 

4-3 

13000 

70.9 

2i" 

6i" 

2  6 

1200   j        13. 1 

225 

1.6" 

8" 

si" 

S 

3.67s 

1 1 06c 

77.2 

1 

3 

8" 

2.667!     900        11.78 

300 

1.6" 

9i" 

7i" 

6.17 

4.  S3 

10500 

73.3 

3^ 

lO-i" 

3 

900  j      13.74 

p   =  pressure  per  square  inch  on  diametrical  plane  of  bearing, 
p   =   7  to  15  pounds  per  square  inch  up  to  55  h.  p. 

=  15  to  23  pounds  per  square  inch  from    55  to  no,  h.  p.  \  for  pinion  bearings. 
=  23  pounds  per  square  inch  from  no  to  300  h.  p 
p    =  1 5  to  40  pounds  on  power  shaft  beanngs. 

Spiral  grooves  in  bearing  to  distribute  the  oil  are  1/16  inch  wide  and  3/64  inch  deep, 
with  holes  at  the  ends  to  watch  the  conditions  of  lubrication. 

Table  39. — Standard  Bearing  Practice  of  DeLaval  Steam  Turbines. 

FORCED  LUBRICATION  FOR  HORIZONTAL  ENGINES 

Fig.  53  is  a  type  plan  of  forced  lubrication  for  reciprocating  engines  of 
battleships  and  armored  cruisers  as  worked  out  by  the  Bureau  of  Steam 
Engineering  of  the  United  States  Navy.  The  location  and  arrangement  of 
tanks,  strainers,  valves,  and  piping  is  shown. 

BALL-AND-SOCKET  SHAFT  BEARINGS 

Fig.  54  shows  the  ring-oiling  line  shafting  box,  and  Fig.  55  a  usual  form 
of  step  bearing  for  vertical  shafts  as  made  by  Wm.  Sellers  &  Company,  Inc. 
The  horizontal  box  may  be  oiled  by  a  syphon  oil  cup  through  the  center  or 


158 


BEARINGS  AND  THEIR  LUBRICATION 


oil  may  be  placed  in  the  reservoir  around  the  top  ball  and  allowed  to  find  its 
way  through  the  small  holes  to  the  shaft.  It  is  customary  to  fill  the  grease 
cups  with  tallow  or  other  grease,  which  will  melt  if  the  box  becomes  at  all  warm. 

LOCOMOTIVE  AND  CAR  BEARINGS 

The  eleven  illustrations,  Figs.   56  to  66,  inclusive,  are  reproduced  from 
designs  of  the  Baldwin  Locomotive  Works.     Two  are  of  locomotive  driving 


rn 


T~l 


U'ln^^^rJ 


7IG.   56. 

LOCO] 

^ 

^> 

->l 

fn:^ 

%  Brass 
Plugs 

. 

K 

-1 

-^>i 

C^D 

LOCOMOTIVE    DRIVING    BOX. 


/^%- 

i 

/iM=n 

:-Hl»A 

pi 

^ 

1 

r 

Hfl 

-f 

II 

-P 

-1-2— Journal- 


T1- 


n 


Wn\ 


Detail  of 
Babbitt. 


FIG.    57.      LOCOMOTIVE    DRIVING    BOX. 

boxes,  two  of  tender  boxes,  one  of  an  engine  truck  box,  one  of  a  crosshead, 
one  of  an  eccentric  strap  and  four  of  rod  stubs.  A  few  dimensions  are  included 
in  each  case  to  give  an  idea  of  proportions. 


TYPICAL   DESIGNS  AND   CONSTRUCTIONS 


159 


As  a  matter  of  interest,  Fig.  67  is  reproduced.  It  is  of  a  ring  oiling  car  box 
invented  and  developed  by  W.  O.  Dunbar,  and  at  one  time  experimented  with 
on  the  Pennsylvania  Railroad. 


For  Journal  BVz  x  10  Inches. 


FIG.    58.      LOCOMOTIVE   TENDER   BOX. 

Via  S'teel  Inserts 


FIG.    59.      L0C0M0TI\rE    TENDER   BOX. 

KINGSBURY  THRUST  BEARING 

Fig.  68  shows  in  partial  elevation  and  section,  a  thrust  bearing  developed 
and  patented  by  Professor  Kingsbury. 


i6o 


BEARINGS  AND  THEIR  LUBRICATION 


The  design  differs  from  the  ordinary  collar  bearing  in  that  one  of  the  pair 
of  collars  is  divided  into  several  sectors  or  '"shoes"  which  are  individually 
supported  on  spherical  seats  whereby  each  shoe  is  enabled  to  adjust  itself  to 


(    yi    I  XT 


-J    II  Ij   Ll 

11-11  = 


mi 

ST     I 


uu. 


-M- 


Vs  Oil  Holes 


1 


2a 


T-i^' 


-Oil-Groov 


iZTTTT 


,       i-M-X—pellaT fj 


FIG.  6o.   LOCOMOTIVE  TRUCK  BOX. 


FIG.  6l.   LOCOMOTIVE  CROSSHEAD. 


the  best  alinement  and  pressure  distribution  at  the  bearing  face.  The  concave 
seats  for  the  shoes  are  formed  in  a  ring  which  has  a  spherical  seat  in  the  housing, 
for  maintaining  group  alinement  of  the  shoes.     Lubrication  is  effected  by  a 


TYPICAL   DESIGNS  AND  CONSTRUCTIONS 


l6l 


bath  of  oil  in  the  case  of  vertical  shafts,  and  by  flood  lubrication  with  horizontal 
shafts. 

The  considerations  leading  to  the  adoption  of  the  construction  are  as  follows: 
Both  theory  and  experiment  have  shown  that  in  shaft  journals  in  which  perfect 
lubrication  is  maintained,  the  film  of  oil  is  invariably  wedge-shaped,  being 


'1'2  Iron  Pipe 


—        i.     I,    XlUll    J.   li^c  I 1^ 


111      Tin 


M 


K^M 

m- 


FIG.    62.      LOCOMOTIVE   ECCENTRIC    STRAP. 


w 

ndleU 


Spi 
Style  No.  1         Spindle 

Style  No.  2 


Oil  Cup 
Style  No.  2 

Oil  Cup 
Style  No.  3 


erg 


t^Jt^j^i 


FIG.    63.      LOCOMOTIVE    ROD    STUB. 


thicker  at  the  edge  where  the  oil  enters  and  thinner  toward  the  leaving  edge. 
In  the  ordinary  collar  bearing,  the  bearing  surfaces  remain  strictly  parallel, 
hence  no  wedge-shaped  film  of  oil  can  form,  and  even  with  bath  lubrication 
the  action  of  the  lubricant  is  imperfect,  thus  the  specific  loads  are  necessarily 
relatively  low  and  the  coefficients  of  friction  high.     The  shoe  construction  on 


l62 


BEARINGS  AND  THEIR  LUBRICATION 


the  other  hand  enables  the  wedge-shaped  films  to  form  and  thus  perfect  lubri- 
cation is  maintained.  The  leading  edge  of  each  shoe  is  depressed  by  the  enter- 
ing oil,  giving  automatically  the  proper  slight  inclination  between  the  bearing 


L'-j-Pj 


ii 


FIG.    64.      LOCOMOTIVE    ROD    STUB. 

/^/32  Bushing  Bronze 


■I 


A_S 


^^n 


nz 


J) 


FIG.    65.      LOCOMOTIVE    ROD    STUB. 

surfaces.     That  this  actually  takes  place  is  shown  by  the  fact  that  such  wear 
as  occurs  on  the  shoes  is  first  indicated  at  the  leaving  edge. 

The  shoe  type  bearings  have  shown  good  results  under  tests  and  in  service. 


TYPICAL  DESIGNS  AND   CONSTRUCTIONS 


163 


Pressures  from  250  to  1000  lb.  per  square  inch  are  borne  under  ordinary  con- 
ditions with  coefficients  of  friction  as  low  as  0.0022.     Test  runs  have  been  made 


FIG.    66.      LOCOMOTIVE   ROD   STUB. 


FIG.   67.      RING-OILING  CAR  BOX. 

at  a  circumferential  speed  of  4500  ft.  per  minute  with  pressures  up  to  2600  lb. 
per  square  inch,  without  causing  failure  of  the  oil  film;  and  at  lower  speeds  a 
pressure  of  5000  lb.  has  been  reached  experimentally. 


164 


BEARINGS  AND  THEIR  LUBRICATION 


TEXTILE  MACHINERY  BEARINGS 

W.  S.  South  wick  sums  up  American  practice  in  the  design  and  application 
of  the  bearings  of  textile  machinery,  thus: 

■  "Among  the  kinds  of  bearings  generally  used  on  textile  machinery,  there  are  two  or  three 
of  interest.  The  bearing  most  generally  used  is  a  simple  cast-iron  box  mating  a  steel  journal. 
Many  of  the  caps  for  larger  sized  bearings  have  a  recess  cast  in  their  upper  side  to  form  an  oil 


^Pr3 


Drain 


\^^_,^^^<^m 


FIG.    68.      KINGSBURY   THRUST   BEARING. 


well;  the  smaller  bearings  simply  have  a  drilled  and  countersunk  oil  hole.     A  good  grade  of 
machinery  oil  is  used  for  lubrication  supplied  by  a  squirt  can. 

"There  are  no  commonly  accepted  factors  for  design.  In  most  cases  the  design  is  governed 
largely  by  experience  and  varies  with  conditions.  A  device  that  has  been  developed  to  a  very 
high  state  of  perfection,  from  a  mechanical  standpoint,  is  the  spindle.  Over  500  patents  have 
been  issued  on  it.  As  now  built,  it  runs  at  from  8000  to  12,000  revolutions  per  minute  ver}- 
smoothly,  with  little  wear,  and  requiring  but  a  small  amount  of  power.  The  blade  which  is 
made  of  hardened  steel,  accurately  ground  to  a  slight  taper,  runs  in  a  cast-iron  bolster  partly 
covered  with  cotton  packing  to  reduce  vibration.  In  some  constructions  the  bolster  is  adjust- 
able to  compensate  for  wear.  The  cast-iron  base  in  which  the  bolster  is  fitted  has  on  one  side  a 
projection  adapted  to  receive  the  oil  and  convey  it  to  the  revolving  parts  which  are  surrounded 
by  oil.     A  light  clean  spindle  oil  is  used,  generally  supplied  by  an  oil  can. 


TYPICAL   DESIGNS  AND  CONSTRUCTIONS  165 

"Large  vertical  twister  rings  are  often  equipped  with  a  self-oiling  device.  This  consists  of 
a  wick  which  lubricates  the  traveler  at  a  point  below  its  contact  with  the  yarn.  The  bearings 
on  the  parts  which  operate  the  harnesses  and  shuttle-boxes  of  fancy  worsted  and  woolen  looms 
are  malleable  iron  and  steel  case-hardened.  These  bearings  are  about  3/4  inch  in  diameter 
and  5/32  inch  wide.  They  are  lubricated  with  an  ordinary  oil  cup.  In  a  few  cases,  chilled 
cast-iron  journals  and  cast-iron  boxes  are  used.     They  are  oiled  in  the  same  manner. 

The  temple  roll  is  an  illustration  of  a  slow  running  bearing  where  a  brass  stud  forms  the 
journal  and  a  wooden  bushing  previously  boiled  in  linseed  oil  the  lining  shell.  This  is  inserted 
into  the  roll.  Albany  grease  is  used  for  lubricating,  being  placed  in  the  bushing  before 
assembling." 


SECTION  X 

Hints  on  the  Care  of  Bearings 

Never  start  a  new  engine  or  machine  without  examining  the  fit  of  the  bear- 
ings and  knowing  it  to  be  right. 

Never  start  up  any  machine  without  knowing  that  its  bearings  are  properly- 
oiled. 

A  good  way  to  adjust  new  engine  main  bearings  is  to  set  the  boxes  up  tight 
and  slack  back  the  adjusting  screws  one  face  of  the  head.  Then  start  the  engine, 
watch  it  carefully  and  tighten  up  the  screws  gradually  until  the  pound  is  all 
taken  out. 

Remember  that  the  clearance  in  a  cold  bearing  is  greater  than  in  the  same 
bearing  at  running  temperature.  When  warming  up  the  journal  grows  larger 
faster  than  the  bore  of  the  bearing. 

Remember  that  after  a  machine  has  been  idle  for  some  time  the  oil  is 
probably  squeezed  out  from  between  the  bearing  surfaces. 

A  reasonable  slackness  for  engine  bearings  from  6  to  12  in.  in  diameter  is 
about  1/200  in.,  for  crank  pin  and  cross-head  pin  bearings  up  to  4  in.  in  diameter 
about  1/250  in.  These  dimensions  can  easily  be  approximated  by  setting  the 
bearing  up  tight,  then  slacking  back  a  certain  number  of  faces  on  the  wedge 
bolt -head  depending  upon  a  number  of  threads  and  the  taper  of  the  wedges. 

A  bearing  set  up  too  slack  is  nearly  as  bad  as  one  too  tight.  Pounding  will 
make  a  box  run  hot  and  is  liable  to  peen  the  brass  or  babbitt  and  destroy  the  fit. 

Don't  squirt  oil  into  an  oil  hole  and  hope  that  it  will  reach  the  journal.  See 
that  the  hole  is  open  and  not  filled  with  dirt. 

Don't  put  the  end  of  the  spout  of  an  oil  can  into  an  oil  hole  and  pump  the 
bottom  unless  there  is  oil  in  the  can. 

Squirt  oil  into  the  oil  holes  and  not  onto  the  floor— in  the  latter  place  it  may 
lubricate  your  shoes  but  it  does  not  help  the  bearings. 

In  cold  weather  a  little  heat  will  help  the  oil  to  run  out  of  your  oil  can — the 
top  of  the  steam-chest  is  a  favorite  place  for  this  warming  up. 

Don't  neglect  a  bearing  because  it  is  under  a  bench,  or  in  a  dark  corner,  or 
in  some  out-of-the-way  place.  "Out  of  sight,  out  of  mind"  is  poor  policy  to 
apply  to  bearings. 

Keep  an  eye  on  the  oil  tanks  and  do  not  let  them  get  too  low.  Good  oil  can- 
not always  be  gotten  in  a  hurry  and  some  bearings  may  suffer, 

166 


HINTS  ON  THE  CARE  OF  BEARINGS  167 

In  general,  the  best  oil  to  use  on  bearings  is  the  lightest  oil  with  which  they 
can  be  run  safely. 

Kerosene  is  successfully  used  as  a  lubricant  for  high-speed,  hardened  steel 
spindles  and  bearings,  were  the  pressures  are  very  light. 

Don't  depend  entirely  on  the  sight  glass  to  tell  you  how  much  oil  is  in  the 
well  of  an  important  bearing. 

When  babbitting  a  box,  a  piece  of  paper  wrapped  around  the  shaft  or 
mandrel  will  keep  the  hot  metal  from  chilling  and  reduce  the  amount  of 
subsequent  scraping. 

If  possible  it  is  a  good  plan  to  heat  a  bearing  shell  before  babbitting. 

Never  pour  babbitt  into  a  shell  that  contains  any  water  or  is  not  perfectly  dry. 
Many  serious  burns  have  come  from  a  failure  to  observe  this  rule. 

If  several  small  bearings  of  the  same  size  are  to  be  babbitted,  it  is  better  to 
use  a  babbitting  arbor  rather  than  a  finished  shaft.  If  the  arbor  is  sprung  a 
little  from  the  heat,  no  harm  is  done. 

For  severe  service  it  is  a  good  plan  to  peen  the  surface  of  a  babbitted  box 
before  it  is  bored  or  reamed  to  size. 

Don't  try  to  run  babbitt  in  bearings  subjected  to  heavy  shock  and  pound. 

Don't  forget  to  cut  oil  grooves  when  you  rebabbitt  a  box. 

Don't  forget  to  drill  the  oil  hole  when  you  rebabbitt  a  box. 

Don't  cut  oil  grooves  through  the  end  of  a  box.     The  oil  will  run  out. 

In  general,  oil  grooves  should  be  cut  at  riglit  angles  to  the  direction  of  motion 
of  the  journal. 

If  oil  will  not  follow  the  grooves  of  a  tightly  fitted  bearing,  try  venting  by  a 
scratch  with  a  file  or  scraper  through  to  the  end  in  order  to  let  out  the  air. 

Be  sure  that  both  ends  of  oil  tubes  are  securely  fastened  so  that  they  cannot 
jar  loose. 

Any  self -oiling  bearing  with  an  oil  well  should  be  washed  out  at  intervals. 
Be  sure  the  well  is  refilled  before  the  machine  is  started. 

Be  sure  that  there  are  no  points  in  oil  tubes  higher  than  the  point  where  oil 
is  introduced.     Oil  will  not  run  up  hill. 

When  making  new  oiling  rings  be  sure  that  they  are  smoothly  finished  and 
without  any  corners  or  edges  to  catch  when  in  use  and  stop  the  rings  from 
revolving. 

When  putting  in  new  chains  for  self-oiling  bearings,  see  that  they  do  not 
drag  on  the  bottom  of  the  oil  well.  If  they  do  hit  the  bottom,  a  hole  will  be 
worn  through  in  time,  letting  out  the  oil  and  probably  ruining  the  bearing. 

Don't  let  induction  motor  bearings  wear  sufficiently  to  drop  the  rotor 
against  the  stator.    Put  in  new  bearings. 

Felt  wipers  should  be  taken  out  at  intervals  and  soaked  in  kerosene  to  soften 
them  and  remove  the  glaze. 


1 68  BEARINGS  AND  THEIR  LUBRICATION 

Don't  let  a  hot  box  begin  to  smoke  before  you  attend  to  it.  You  should 
smell  it  long  before  you  see  it. 

The  rapid  heating  of  a  bearing  is  a  danger  sign.  If  a  bearing  takes  two  or 
three  hours  to  reach  a  temperature  uncomfortably  hot  for  the  hand,  it  is  prob- 
ably safe,  but  if  that  same  temperature  is  reached  in  lo  or  15  minutes  there  is 
liable  to  be  trouble. 

If  you  have  a  hot  box  on  a  main  engine  bearing,  first  loosen  it  up  a  little  all 
around  and  then  give  it  a  generous  dose  of  cylinder  oil  through  the  feel  holes, 
but  don't  try  to  feed  it  through  the  ordinary  oil  channels;  they  are  too  small.  If 
this  oil  will  stick  to  the  shaft  it  is  apt  to  cure  the  trouble. 

Another  good  remedy  is  a  mixture  of  graphite  and  oil — one  part  of  graphite 
to  10  of  oil.  But  do  not  feed  this  until  the  graphite  has  been  thoroughly  wetted. 
Dry  graphite  will  not  help  a  hot  box. 

If  there  is  no  graphite  handy  use  flowers  of  sulphur  in  the  same  way;  or 
failing  in  that,  try  ground  talc.  These  should  be  mixed  with  oil  in  the  same 
proportions,  one  of  the  powder  to  10  of  oil,  and  the  dry  part  of  the  mixture  must 
be  thoroughly  wetted  before  it  is  applied. 

In  aggravated  cases  a  stream  of  water  from  a  hose  will  accomplish  a  good 
deal  in  bringing  down  bearings  temperatures,  or  ice  may  be  used  packed  around 
the  bearings. 

Hot  crank  pin  boxes  may  be  cooled  by  fixing  a  hose  in  such  a  position  that  the 
crank  will  strike  through  the  running  water  at  about  the  upper  part  of  its  swing. 
The  same  scheme  may  be  used  with  hot  cross- head  pin  boxes. 

Small  bearings  that  heat  can  sometimes  be  helped  by  flooding  with  kerosene 
or  gasoline  to  wash  out  any  dirt,  grit  or  metal  particles  that  may  be  in  them. 
However,  this  method  must  be  used  with  caution  for  the  lubricating  oil  is  also 
washed  away.  A  large  supply  of  regular  lubricating  oil  must  be  at  once  poured 
in.  If  on  hand  the  kerosene  should  be  used  in  preference  to  gasoline  as  the 
latter  cuts  the  lubricating  oil  more  quickly. 

Sometimes  a  small  bearing  that  heats  can  be  helped  by  feeding  all  of  the  oil 
that  can  be  run  through  it — flooding.  This  tends  to  bring  down  the  temper- 
ature and  may  wash  away  some  foreign  particles. 

If  a  babbitted  bearing  heats  pull  it  to  pieces  and  look  for  black  spots. 
Those  are  the  places  to  scrape. 

Bearings  can  often  be  run  much  hotter  than  is  generally  thought.  A  bearing 
hot  enough  to  blister  the  hand  will  run  without  injury. 

All  important  bearings,  even  of  the  self-oiling  type,  should  be  examined  at 
frequent  intervals.  In  many  shops  it  is  the  practice  to  inspect  all  motor  and 
generator  bearings,  at  least  once  each  day,  and  to  thoroughly  overhaul  all  of  the 
line  shaft  bearings  either  once  every  six  months  or  once  a  year,  depending 


HINTS  ON  THE  CARE  OF  BEARINGS  169 

upon  the  nature  of  work.  This  means  taking  the  bearings  to  pieces,  thoroughly 
washing  all  parts,  cleaning  out  the  oil  wells  and  refilling  with  fresh  oil. 

Where  large  quantities  of  lubricating  oil  are  used  an  oil  filter  will  save  money. 

As  far  as  possible,  do  not  let  any  oil  from  your  bearings  escape  either  by 
dripping  or  through  atomizing.  This  is  one  of  the  most  difficult  points  to  care 
for  in  practice,  particularly  with  high  speed  machinery.  If  oil  is  thrown  out  in 
quantities  as  in  large  drops  or  small  streams,  it  is  comparatively  easy  to  stop  it. 
Felt  wiping  rings  at  the  end  of  the  box  are  usually  effectual  but  the  material  must 
not  be  too  porous,  or  it  may  act  as  a  sucking  medium  and  instead  of  being  a 
remedy  be  anoiher  source  of  trouble. 

It  is  not  unusual  to  fmd  a  central  station  where  the  walls  of  the  room  are 
covered  with  a  coating  of  oil  and  dirt;  around  the  generators  and  motors  the 
atomized  oil  gathers  particles  of  dust,  settles  in  the  air  passages  and  may 
eventually  shut  off  the  ventilation.  It  is  most  difficult  to  stop  this  atomizing 
action.  What  must  be  done  is  to  prevent  any  current  of  air  passing  through 
the  bearing  itself  or  across  the  surface  of  any  body  of  oil.  If  this  takes  place 
oil  will  be  vaporized  and  diffused  throughout  the  entire  room.  The  points 
where  oil  escapes  can  usually  be  determined  by  pieces  of  white  paper  or  pieces 
of  sheet  steel  painted  a  light  color.  The  oil  stains  will  show.  The  only  remedy 
is  to  entirely  shut  off  all  air  currents  so  that  they  cannot  reach  the  oil.  At 
the  same  time  this  is  often  a  baffling  problem  to  solve  successfully. 


PART  11. 

BEARINGS  WITH  ROLLING  CONTACT 


SECTION  I 

Rolling  Friction  and  Factors  or  Design  for  Ball  Bearings 

The  exact  nature  of  rolling  friction  is  not  accurately  known,  but  much 
splendid  experimental  work  has  been  done  with  bearings  that  involve  this  princi- 
ple and  we  have  empirical  laws  to  govern  designs.  However,  there  is  no  theore- 
tical investigation  giving  results  that  can  be  applied  to  ball  and  roller  bearings 
in  the  way  that  the  work  of  Professor  Osborne  Reynolds  can  be  applied  to 
journal  bearings  with  sliding  contact. 

A  consideration  of  the  properties  of  ordinary  journal  bearings — bearings 
with  sliding  contacts — and  ball  and  roller  bearings  presents  a  series  of  contrasts. 

At  the  outset  we  have  sliding  contact  contrasted  with  rolling  contact. 

In  the  journal  bearing  the  friction  is  governed  by  the  quality  of  the  lubrica- 
tion. The  better  the  lubrication  the  less  the  friction.  In  ball  and  roller  bear- 
ings the  reverse  is  true.  It  has  been  demonstrated  by  experiment  that  a  well 
designed  ball  or  roller  bearing  may  give  a  lower  coefficient  friction  when  run- 
ning clean  and  dry  than  when  lubricated  with  even  a  thin  oil.  This  must  not 
be  taken  to  mean  that  such  bearings  should  be  run  without  oil  or  grease,  for 
the  reverse  is  necessary.   ' 

In  perfectly  lubricated  journal  bearings  there  is  no  metallic  contact  between 
the  surfaces.     In  ball  and   roller   bearings   there  is  always   metallic  contact. 

In  journal  bearings  many  different  kinds  of  materials  are  used,  running  from 
wood  and  soft  babbitt  to  hardened  steel.  With  a  few  rare  exceptions,  the  only 
material  used  for  ball  and  roller  bearings  is  steel,  and  in  general  it  is  hardened. 

The  coefficient  of  friction  for  sliding  contact  is  much  higher  than  for  rolling 
contact.  The  coefficient  of  rest  and  slow  motion  for  sliding  contact  is  very 
much  greater  than  for  motion  at  ordinary  speeds.  The  coefficient  of  friction  for 
the  beginning  of  motion  with  rolling  contact  is  not  much,  if  any,  greater  than 
for  motion  at  ordinary  speeds.  For  this  reason  ball  and  roller  bearings  are 
peculiarly  adapted  for  machinery  having  a  high  starting  torque. 

In  journal  bearings  running  at  a  high  rubbing  speed  the  dissipation  of  heat 
becomes  an  important  problem.  With  properly  designed  and  operated  ball 
and  roller  bearings  the  amount  of  energy  liberated  in  heat  is  so  slight  that  it  is  a 
negligible  factor  in  design.     Cooling  devices  are  not  used. 

With  journal  bearings  practice  indicates  that  a  limiting  speed  is  soon  reached. 
For  illustration,  internal  grinding  fixtures  with  such  bearings  are  seldom  run 

173 


174  BEARINGS  AND  THEIR  LUBRICATION 

over  12,000  to  15,000  revolutions  per  minute.  By  contrast,  internal  grinding 
spindles  mounted  on  ball  bearings  are  commonly  run  at  speeds  of  30,000  revolu- 
tions per  minute,  are  reported  to  have  been  run  commercially  at  a  speed  of 
100,000  revolutions  per  minute,'  and  experimentally  at  a  speed  of  120,000  revo- 
lutions per  minute. 

For  heavy  loads  supported  on  vertical  step  bearings  the  ordinary  step  type 
with  sliding  contact  has  frequently  been  a  source  of  trouble.  On  the  other 
hand,  there  are  ball  and  roller  thrust  bearings  in  use  carrying  loads  up  to 
400,000  lb.  and  in  one  installation  a  load  of  over  2,000,000  lb.  is  sustained  on  a 
plain  roller  thrust  bearing. 

FORMULAS  FOR  UNIT  LOADS 

The  formulas  for  specific  load,  or  load  per  unit  of  carrying  element,  are  as 
follows: 

W 
p= for  roller  journal  bearings. 

""id 
c 

W 
p= for  radial  ball  bearings. 

c 

W 

p=  —  for  roller  and  ball  thrust  bearings. 
n 

In  these  formulas  Pr=the  total  load,  />=the  unit  load,  or  load  per  carrying 
element,  Z=the  length  of  the  journal  bearings,  c^=the  diamenter  of  the  rollers 
in  a  roller  bearing,  Z)  =  the  diameter  of  the  balls  in  a  ball  bearing,  w=the 
number  of  rollers  or  balls,  as  the  case  may  be  and  c=a  constant  =  5  for 
conditions  recited  below. 

It  will  be  noted  that  the  entire  number  of  balls  or  rollers  is  not  considered 
as  carrying  load,  but  the  total  number  divided  by  c.  It  is  clear  that  not  more 
than  half  of  the  balls  (or  rollers)  can  be  in  the  loaded  side  of  the  journal.  It  is 
also  clear  that  that  ball  immediately  in  line  with  the  direction  of  the  load  will 
carry  most,  while  the  balls  (or  rollers)  at  either  side  will  carry  consecutively  less 
and  less.  For  bearings  that  would  take  between  15  and  30  balls  or  rollers  it 
will  do  to  take  ^=5.  For  other  conditions  the  proportionate  load  imposed  on 
each  ball  (or  roller)  must  be  figured;  this  is  readily  done  mathematically  or 
graphically.  In  many  forms  of  ball  and  roller  bearings  the  balls  (or  rollers) 
are  separated,  thus  their  number  is  less  than  the  possible  maximum  that  might 
be  inserted  did  they  lie  in  actual  contact;  n  in  the  above  formula  is  to  be  taken 
always  as  this  possible  maximum. 


ROLLING  FRICTION  AND  FACTORS  OF  DESIGN 


175 


Professor  Stribeck  gives  values  of  p  for  roller  bearings  and  continuous  ser- 
vice from  85  to  155  lb. 

Fig.  69  is  taken  from  Stribeck's  experiments  and  shows  the  relationship 


.uou 

11 

i 

i 

.040 

1 1 

1! 

I 
1 
1 

i 

i 

i 

.030 

|! 

11 

1 

il 

ll 
11 

\ 

"i! 

.020 

\ 

■Mil  \ 
MM 

\ 

1  l\V 

\ 

\ 

D 

1 

\ 

\ 

4\ 

\ 

\ 

.010 

\v\ 

\ 

\ 

"-. 

1^, 

\^ 

\ 

s^ 

•-^ 

^^^ 

1' 

% 

^''- 



Sx 

W 

E^^^ 

h-: 

^2 

\ 

^-0. 
^^^ 

■0— T*-*^ 

^ 

tS_ 

— 

S:,_ 

-^ 

— IP— 

s:~ 

10 


20  30  40  50  60  70  80 

Specific  Load  in  Kilograms,  (Kilograms  and  C.G.S.  Units) 


90 


100 


FIG.    69.      RELATIONS  OF  COEFFICIENTS  OF  FRICTION  FOR  PLAIN,  ROLLER  AND  BALL 

BEARINGS  (Stribeck). 


between  the  coefficient  of  friction  and  unit  load  for  journal  bearings,  several 
types  of  roller  bearings  and  one  type  of  ball  bearing.  Referring  to  the  specific 
curves:     Curves  S  are  for  a  babbitted  journal — S^  at  iioo  revolutions  per 


176  BEARINGS  AND  THEIR  LUBRICATION 

minute,  6*2  at  380,  ^3  at  64  and  S^  at  12.  Curves  A  are  for  rollers  in  a  bronze 
cage,  circumferentially  fitting  and  supporting  each  roller  throughout  its  length. 
A  sleeve  on  the  shaft  and  a  liner  in  the  box  carry  the  rolls.  Hard  and  soft 
rollers  were  tested;  A,  at  11 00  revolutions  per  minute,  A2  Sit  285,  ^3  at  190. 
Curves  B  are  for  hollow  rollers  on  loosely  fitted  pins  that  connect  the  end  cages. 
No  liners  are  used  on  the  shaft  nor  in  the  box.  The  speeds  are  190  to  760  revo- 
lutions per  minute.  Curves  D  are  for  hollow,  flexible,  short,  sheet  steel  rollers 
on  loosely  fitted  pins  that  connect  the  end  cages.  No  liners  were  used  on  the 
shaft  or  in  the  box.  The  speed  was  iioo  revolutions  per  minute.  Curves  E 
are  for  flexible  spiral  rollers,  placed  loosely  in  the  box  and  alined  by  three  longi- 
tudinal bars  connecting  the  end  cages.  No  liner  for  the  shaft  or  box  are  em- 
ployed. The  speeds  are  56  to  iioo  revolutions  per  minute.  Curve  F  is  for  a 
ball  bearing  of  the  two-point,  curved  ball-race  type  (with  race  curvature  but 
slightly  in  excess  of  ball  curvature)  at  65  to  11 50  revolutions  per  minute.  The 
range  of  permissible  specific  load,  as  indicated  by  the  appearance  of  the  bearing 
after  long  continued  runs,  is  shown  by  smafl  circles  on  the  curves. 

The  above  curves  and  description  are  taken  from  a  translation  and  amplifi- 
cation of  the  work  of  Professor  Stribeck,  made  by  Henry  Hess,  and  found  in  the 
Trans.  A.S.  M.  E.,  Vol.  XXVII.  That  celebrated  work  lays  the  foundation  for 
the  intelligent  design  of  ball  and  roller  bearings. 

As  ball  and  roller  bearings  are  made  by  only  a  few  firms  and  as  many  of 
the  devices  are  patented  it  will  be  necessary  to  refer  frequently  to  specific  bear- 
ings in  the  following  pages. 

PROFESSOR  STRIBECK'S  INVESTIGATION 

Ball  bearings  first  came  into  general  use  in  the  bicycle.  They  were  of  the 
so-called  cone-bearing  type  and  in  general  had  an  adjustable  cup  and  cone  and 
two-point  contact.  They  were  designed  by  rule  of  thumb  and  the  basic 
principles  were  in  obscurity. 

This  was  the  condition  of  affairs  in  1898,  when  the  problem  of  determining 
the  carrying  capacity  of  balls  and  the  laws  for  the  design  of  ball  bearings  was 
attacked  by  Professor  Stribeck,  then  director  of  the  Scientific  Technical  Experi- 
mental Laboratories  at  Neu-Bablesberg,  Germany.  His  solution  set  a 
thorough  precedent. 

He  first  determined  the  load  sustaining  capacity  of  balls  and  found  that 
it  is  proportional  to  the  square  of  the  ball  diameter,  and  very  much  dependent 
upon  the  shape  of  the  surfaces  or  grooves  between  which  the  ball  is  held. 

CURVATURE  OF  THE  RACE  GROOVES 

The  governing  principle  in  regard  to  the  shape  of  the  races  is  that  the  ball 
shall  roll  over  them  and  not  slide.  '  Theoretically  balls  touch  their  contact  sur- 


ROLLING  FRICTION  AND  FACTORS  OF  DESIGN  177 

faces  in  points  only.  Practically,  even  with  balls  of  hardened  steel  of  high 
elasticity  they  touch  over  surfaces  of  appreciable  area  instead  of  at  points. 
Theory  and  experiments  both  prove  that  the  two-point  contact  is  the  best.  To- 
day all  of  the  better  known  ball  bearings  and  most  of  the  ball-thrusts  have 
two-point  contact. 

If  it  were  possible  to  so  shape  the  groove  that  all  parts  would  sustain  load 
equally  we  would  have  a  condition  of  maximum  sustaining  capacity  of  the  ball. 
Such  a  groove  would  be  represented  by  a  profile  semi-circular  in  section  with 
radius  of  curvature  equal  to  one-half  the  ball  diameter.  Because  of  excessive 
friction  this  form  is  not  practicable.  It  is  evidently  one  limit  of  the  curvature  of 
the  race. 

Turning  to  the  other  limit,  if  we  begin  with  a  cylindrical  race  the  sustaining 
power  increases  without  a  proportinate  increase  of  friction  as  the  curvature  is  in- 
creased up  to  the  point  of  equality  of  curvature,  when  sliding  sets  in  and 
with  that  a  change  from  rolling  to  sliding  friction.  The  best  results  come 
from  proportions  in  which  ^1  =  9/16  D  and  ^2=  25/48  D.  Here  ri  =  the 
radius  of  curvature  of  the  outer  race  groove,  r2  =  the  radius  of  curvature  of  the 
inner  race  groove,  and  Z)  =  the  ball  diameter.  The  radius  for  the  outer  groove 
will  be  seen  to  be  slightly  greater  than  the  other,  the  reason  for  this  being  that 
the  concavity  of  the  outer  race  tends  to  embrace  more  of  the  ball  than  the  inner 
race,  which  is  convex.  Even  greater  approximation  to  equality  of  race  and  ball 
curvature  is  desirable,  but  is  impracticable  owing  to  workshop  difficulties. 

LOAD  CARRYING  CAPACITY  OF  RADIAL  BEARINGS 

Early  investigators  of  ball  bearings  took  into  consideration  the  crushing 
strength  of  the  balls.  This  is  an  erroneous  starting-point.  The  capacity  of 
a  ball  in  a  ball  bearing  is  not  its  crushing  strength  by  any  means,  but  it  is  its 
resistance  to  permanent  deformation;  some  deformations  take  place  with  any 
appreciable  load  no  matter  how  high  grade  the  material  from  which  the  balls  are 
made.  The  load  carrying  capacity  of  a  ball  bearing  is  directiy  proportional  to 
the  number  of  balls  and  to  the  square  of  the  ball  diameter.  According  to 
Stribeck  this  capacity  P  is  expressed  by  the  equation  P=k  n  U^  in  which  n= 
the  number  of  balls,  D  the  ball  diameter,  and  k  a  cofficient  depending  upon 
the  type  of  bearing,  the  material  from  which  it  is  made  and  the  speed  at  which 
it  is  run. 

Mr.  Hess  gives  values  of  this  factor  k  for  a  special  steel  used  by  his  firm  as 
follows,  provided  that  D  is  expressed  in  units  of  an  eighth  of  an  inch: 

For  uninterrupted  race  grooves  with  cross-sectional  curvature  as  mentioned 
above,  separated  balls,  uniformly  distributed  load  and  a  uniform  speed  up  to 
3,000  revolutions  per  minute,  ^  =  9. 


178  BEARINGS  AND  THEIR  LUBRICATION 

For  full  type  bearings  with  the  filling  opening  in  one  race  at  the  unloaded 
side,  otherwise  as  above,  ^  =  5. 

For  both  ball  tracks  interrupted  by  filling  openings,  inelastic  cage  separators 
for  the  balls,  or  for  full  ball  type;  and  speeds  not  over  2,000  revolutions  per 
minute  with  a  uniform  distributed  load,  k  =2.5. 

For  thrust  load  on  a  radial  bearing  of  the  first  type  in  this  tabulation  k  =0.9. 

In  general,  the  larger  the  number  of  balls,  the  smaller  the  value  of  k.  The 
tadial  load  bearing  is  up  to  high  speeds,  practically  unaffected  by  the  speed  as 
ro  its  carrying  capacity. 

LOAD  CARRYING  CAPACITY  OF  THRUST  BEARINGS 

Ball  thrust  bearings  are  made  in  three  general  types.  In  the  first  both 
races  are  flat,  in  the  second  one  race  is  flat,  the  other  grooved.  In  the  third 
both  races  are  grooved.     The  load  carrying  equation  given  by  Mr.  Hess  is: 

P=      — 

Vs 

In  which  F=  the  load  capacity  in  pounds,  k^  =  Si  factor  determined  by  exper- 
iment and  varying  with  the  material  and  the  shape  of  the  ball  track,  n=the 
number  of  balls,  D=the  ball  diameter  and  5=  the  number  of  revolutions  per 
minute  of  the  bearing.  Up  to  about  3000  revolutions  per  minute  the  effect  of 
speed  varies  with  its  cube  root. 

Mr.  Hess  gives  the  following  values  for  k^. 

For  material  such  as  used  by  his  firm  and  for  races  having  grooves  with  a 
cross-sectional  radius  equal  approximately  to  0.82  D,  k^  =  2^  to  40. 

For  unhardened  steel,  such  as  is  occasionally  used  for  very  large  races  and 
where  there  is  no  hammering  or  no  sharp  blows,  ^^  =  0.5. 

When  one  or  both  races  are  flat  k^  should  be  reduced  to  one-fourth  the 
above  value. 

The  Standard  Roller  Bearing  Company  gives  the  following  load  capacity 
formulas  for  ball  thrust  bearings  having  a  groove  in  each  washer,  stating  that 
the  ratings  obtained  from  their  use  are  very  conservative  and  give  a  condition 
of  loading  under  which  a  bearing  can  be  guaranteed  if  properly  installed  and 
lubricated. 

The  formula  for  the  light  type  bearing  with  17  balls  is: 

P=32,ooo  : 

in  which  P  =  the  load  capacity  in  pounds, 
D=the  ball  diameter  in  inches, 
n  =the  number  of  balls, 

a  =  the  pitch  diameter  of  the  ball  grooves  in  inches, 
S  =  the  speed  of  the  shaft  in  revolutions  per  minute. 


ROLLING  FRICTION  AND  FACTORS  OF  DESIGN  179 

The  formula  for  the  medium  and  heavy  bearings  having  11  and  9  balls 
respectively  is: 

P=  10, 200       /- 
a's/S 

The  notation  is  the  same  as  given  above. 

Ball  thrust  bearings  are  commonly  of  the  two-point  type  to  which  all  of  the 
preceding  formulas  apply.  However,  a  line  of  thrust  bearings  having  four 
points  of  contact  is  made  by  the  Auburn  Ball  Bearing  Company. 


SECTION  II 

CONSTRUCTION  OF  BALL  BEARINGS 

From  the  fact  that  the  load  carrying  capacity  of  a  ball  bearing  is  directly 
proportional  to  the  number  of  balls,  early  designs  were  of  the  full  type  with  an 
opening  to  admit  the  balls.  Fig.  70  shows  an  entering  notch  at  the  side  cutting 
through  into  the  groove  in  the  upper  race.  This  notch  was  closed  by  means  of 
a  block,  as  shown  at  the  left  in  Fig.  71.     A  modification  of  the  same  idea  was 


FIG.    70.       RADIAL  BALL  BEARING  WITH  BALL  ENTERING  NOTCH  IN  UPPER  RACE. 

a  hole  through  the  outer  race  closed  by  a  screw  plug  as  shown  at  the  right  in 
Fig.  71.  In  both  of  these  constructions  the  closing  piece  was  located  at  the 
point  of  least  pressure.  A  modification  of  the  first  device  used  in  small  bearings 
consisted  in  stopping  up  the  notch  at  the  side  by  means  of  a  small  screw.  These 
early  devices  are  credited  to  Riebe,  Bierschenk  and  Bruhl. 


FIG.    71.       DEVICES    FOR   CLOSING    THE    ENTERING    NOTCH   IN    BALL-BEARING    RACES. 

All  of  these  constructions  reduce  the  load  carrying  capacity  of  the  bearing 
because  of  the  weakened  race  section  where  the  notch  or  opening  is  made  and 
because  after  slight  wear  there  develops  an  interference  between  the  balls  and 
the  edges  of  the  opening.  Furthermore  this  type  of  bearing  cannot  take  any 
thrust  load,  and  in  practically  all  constructions  radial  bearings  are  required  to 

180 


CONSTRUCTION  OF  BALL  BEARINGS  l8l 

take  a  limited  amount  of  thrust  because  of  the  end  play  of  the  shaits  that  they 
support. 

To  overcome  these  obvious  defects  the  partly  filled  type  was  developed. 
Here  the  balls  are  separated  by  means  of  spacing  blocks  or  cages.  As  it  is 
desirable  to  enter  as  many  balls  as  possible  two  general  constructions  have 
grown  up.  The  first  is  known  as  the  eccentrically  filled  bearing,  in  which  the 
inner  race  is  brought  in  contact  with  the  outer  race,  leaving  a  crescent  shaped 
opening  between  them  into  which  all  of  the  balls  but  one  are  entered.  They 
are  then  rolled  around  until  the  outer  and  inner  races  are  nearly  concentric  and 
the  last  ball  is  snapped  into  place  under  a  slight  pressure.  The  races  are  per- 
fectly continuous  and  there  is  nothing  structurally  to  interfere  with  the  free, 
perfect  rolling  of  the  balls.     This  is  the  D.  W.  F.  and  Hess-Bright  Construction. 

HofTmann  also  patented  a  process  in  which  the  outer  ring  is  heated  to  expand 
it,  or  the  inner  cooled  to  contract  it,  or  both  operations  performed  simultane- 
ously. With  this  process  it  is  not  possible  to  introduce  more  than  two  or  at 
the  most  three  more  balls  than  if  no  such  device  is  resorted  to. 

The  other  type  is  made  with  two  modifications.  In  the  first  there  are  two 
filling  notches,  one  in  the  outer  and  one  in  the  inner  race,  sometimes  introduced 
at  an  angle  with  the  circumference,  in  which  case  they  are  inclined  in  opposite 
directions.  The  balls  are  filled  through  these  notches.  The  other  modification 
consists  in  grinding,  on  one  side,  the  inner  face  of  the  outer  ring  eccentric  with 
its  circumference,  so  that  this  ground  surface  is  tangent,  or  nearly  tangent,  to 
the  bottom  of  the  groove.  This  forms  an  opening  through  which  the  balls 
can  be  entered.  The  first  construction  is  used  by  Fichtel  and  Sachs,  the  second 
by  the  Rhinelander  Machine  Works. 

THE  BALL  CAGE 

With  the  use  of  the  pardy  filled  type  it  became  necessary  to  hold  the  balls 
in  positions  equally  spaced  around  the  race.  In  fact,  it  is  practically  impossible 
so  to  proportion  races  and  balls,  even  of  the  full-ball  type,  that  the  latter  exactly 
fit  the  circumference.  Furthermore,  the  cage  has  two  other  important  uses. 
The  one  to  reduce  noise,  the  second  to  prevent  wear. 

If  a  full  ball  type  of  bearing  is  set  in  motion  under  the  usual  vertical  load- 
ing, whenever  a  ball  passes  from  the  region  of  load  to  the  region  where  it  is 
free  from  pressure  it  is  snapped  forward  with  a  motion  analogous  to  the  snap- 
ping of  an  apple  seed  from  between  a  boy's  fingers.  It  strikes  forcibly  against 
the  adjacent  ball  and  inner  race,  bounds  back  against  the  outer  race  and  contin- 
ues to  vibrate.  Before  it  comes  to  rest  it  is  struck  by  the  following  ball  which 
is  then  being  subjected  to  the  same  action.  This  is  a  chief  cause  of  noise  in 
ball  bearings. 


1 82  BEARINGS  AND  THEIR  LUBRICATION 

The  second  trouble  that  the  cage  overcomes  is  wear  of  one  ball  on  another. 
With  full  type  bearings,  it  is  not  uncommon  to  find  balls  each  of  which  has  a 
groove  worn  around  its  circumference  with  a  radius  of  curvature  equal  to  that 
of  the  ball  diameter.  One  cause  of  this  is  the  deflection  of  the  shaft  or  journal, 
as  in  the  case  of  a  large  gear  which  bends  its  shaft  when  its  normal  load  or  an 
overload  is  suddenly  applied.  The  inner  race  of  the  bearing  is  turned  obliquely 
with  reference  to  the  outer  race. 

If  the  axis  of  the  outer  and  inner  races  are  co-incident,  the  path  of  the  ball 
is  a  true  circle.  If,  however,  we  twist  one  race  into  an  oblique  position  with 
reference  to  the  other  and  plot  the  ball  path,  we  will  find  that  it  is  no  longer  a 
circle,  but  a  compound  curve  with  two  cusps.  Thus  in  one  revolution  of  such 
a  bearing,  the  balls  approach  and  recede  from  each  other  twice.  This  action 
interferes  with  the  rolling  action  and  causes  sliding  and  pressure  between  balls 
and  balls  in  full  bearings;  and  between  balls  and  cages  where  the  latter  are  used. 
As  the  cages  are  stationary  with  reference  to  the  balls  and  in  the  best  forms  also 
yield  slightly  this  "binding"  tendency  is  practically  eliminated. 

It  is  important  to  keep  all  ball  bearings  properly  alined. 

Cages  were  devised  to  keep  the  balls  properly  separated  in  a  partly  filled 
tjrpe  bearing,  to  reduce  noise  and  to  prevent  certain  forms  of  ball  wear.  One 
of  the  most  perfect  theoretical  forms  is  that  of  Conrad  which  consisted  in 
separating  the  balls  by  a  short  helical  spring  with  suitable  bearing  plates. 
This  form  was  generally  used  for  several  years  by  the  Hess-Bright  and  D. 
W.  F.  Companies  and  is  still  used  as  a  special  form  for  certain  conditions. 

Other  satisfactory  forms  consist  of  a  cage  made  of  brass  or  some  form  of 
bearing  metal  and  so  constructed  as  to  be  somewhat  elastic.  Many  ball  bearing 
failures  can  be  traced  to  improperly  designed  and  constructed  cages,  in  fact  it 
is  not  unusual  in  a  failed  bearing  to  find  a  cage  literally  ground  to  pieces. 
It  is  not  a  difficult  matter  to  construct  a  cage  which  shall  be  satisfactory  for 
a  given  set  of  conditions,  but  ball  bearings  are  nowadays  manufactured  for 
stock  in  many  thousands  per  day;  it  is  therefore  impossible  to  know  whether  a 
given  bearing  will  be  used  at  a  few  or  many  thousands  of  revolutions  per 
minute,  under  uniform  or  variable  loads  or  speeds,  under  no  or  great  shock,  etc. 
No  cage  has  yet  been  found  that  will  respond  with  equal  satisfaction  to  all  of 
these  many  conditions. 

Many  attempts  have  been  made  to  introduce  a  rolling  member  between 
the  balls  of  ball  bearings  and  the  rollers  of  roller  bearings.  These  are  used 
instead  of  cages  and  are  based  on  the  conception  that  there  will  thus  be  only 
rolling  contact  between  the  various  rolling  elements  and  that  pressure  between 
them  will  then  be  rendered  harmless.  As  a  matter  of  fact  there  can  be  no 
pressure  contact  between  balls  or  balls  and  separators  in  a  correctly  acting 
bearing.     Such  pressure  is  due  either  to  faulty  original  design  or  to  faulty  action 


CONSTRUCTION  OF  BALL  BEARINGS 


183 


from  deflections,  etc.     In  either  case  spinning  or  sliding  will  be  caused  and 
rolling  prevented.     It  is  clear  that  an  element  that  must  roll  to  fulfil  its  mission 
cannot  do  so  when  not  allowed  to  roll  at  the  critical  periods. 
Both  balls  and  rollers  have  been  used  for  this  purpose. 

FRICTION  OF  RADIAL  BALL  BEARINGS 

A  great  deal  has  been  written  in  regard  to  the  smallness  of  the  coefficient 
of  friction  in  ball  and  roller  bearings.  Undoubtedly  this  feature  has  been 
given  much  more  importance  than  it  deserves.  No  argument  or  reiteration  is 
necessary  to  prove  that  bearings  with  rolling  contact  have  less  friction  than 
bearings  with  sliding  contact. 

It  is  impossible  to  generalize  in  regard  to  such  coefficients  of  friction 
In  a  radial  bearing  the  nature  of  the  loading  has  a  direct  influence  upon  the 
shape  of  the  friction  curve.    Figure  72  reproduced  from  the  American  Machinist 


■ 

i 

1.5 

1.4 

1.3 

1.2 

^1.1 

c  1.0 

> 

N 

™% 

^s.oV 

.&>"1.^ 

/ 

X^"^^ 

.2  0.5 

>*"" 

' 

^^ 

c" 

i 

f^ 

k 

- 

- 

^ 

ri^"" 

S' 

f\ 

. 

L 

>a 

r 

^ 

piwau 

n    uai, 

0.4 
0.3 

, 

0.2 

— , 

, 

-Varfab" 

e 

^ 

Radial 

.Ji^ 

— 

— 

1 — 

0.1 

1 

1    1    1" 

1  1  r 

1    1 

- 

400   800 


1200   1600   2000   2400   2800   3200 
Speed  in  R.P.M. 


FIG.    72.      TYPICAL  CURVES    FOR  A  NO.  308    RADIAL    BALL  BEARING  UNDER  VARIOUS  CONDITIONS 

OF    LOADING. 

of  March  4,  1909,  shows  three  curves.  The  first  is  for  a  variable  radial 
load,  the  second  for  a  variable  thrust  load,  the  third  for  a  combination  of  con- 
stant radial  load  and  variable  thrust  load  all  on  a  radial  ball  bearing.  Tables 
40,  41,  and  42  show  the  logs  of  the  tests  from  which  the  curves  were  plotted. 
These  tests  were  made  on  a  special  ball  bearing  testing  machine,  designed  by 
Mr.  Hess  and  built  by  Riehle  Brothers  Testing  Machine  Company.  These 
coefficients  and  curves  are  characteristic  of  this  type  of  bearing. 

Under  a  constant  radial  load  the  coefficient  of  friction  for  a  given  speed  is 
nearly  constant.  With  a  continued  constant  radial  load  and  variable  thrust 
load,  the  coefficient  increases  in  value  and  the  curve  rises  rayidly.  This  con- 
dition is  caused  by  a  sliding  action  which  increases  in  degree  with  an  increase 
in  the  thrust  load.  With  a  variable  thrust  load  only,  the  curve  starts  high, 
drops  sharply  and  then  continues  at  nearly  a  constant  value. 


i84 


BEARINGS  AND  THEIR  LUBRICATION 


It  must  not  be  argued  from  this  one  set  of  curves  that  radial  bearings  should 
not  be  used  under  thrust.  For  speeds  roughly  below  1500  revolutions  per 
minute  the  thrust  type  of  bearing  is  best;  at  about  this  speed  (the  exact  point 
varies  with  changes  in  size)  the  curve  of  the  radial  type  under  thrust  crosses  and 
falls  below  that  of  the  pure  thrust  type. 

It  will  thus  be  good  engineering  practice  to  use  a  single  radial  bearing  to 
carry  both  radial  and  thrust  load,  or  to  use  one  radial  bearing  to  carry  radial 
load  and  a  second  one  to  carry  thrust  load  only,  which  it  is  made  to  do  by  free- 
ing its  outer  race  circumferential! y  and  allowing  it  to  rest  against  its  radial 
faces  only,  or  to  use  instead  of  this  second  radial  bearing  a  separate  bearing 
of  the  pure  thrust  type. 


Radial 

Coefficient  of 

Oil  tempera- 

load 

Vernier 

Vernier 

Difference 

Average 

friction  in 

ture,  degrees 

lb. 

per  cent. 

Fahrenheit 

400 

44875 

44661 

0.00214 

0.00107 

0.1633 

600 

44851 

44652 

0.00199 

0.000995 

0.1517 

800 

44663 

44489 

0.00174 

0.000870 

0-1325 

79.1 

1000 

44581 

44390 

0.0017 1 

0.000855 

0.1302 

1200 

44602 

44439 

0.00163 

0.000815 

0. 1240 

1400 

44524 

44366 

0.00158 

0.00079 

0.1200 

1600 

44506 

44349 

0.00157 

0.000785 

0.1195 

82.5 

1800 

44443 

44276 

0.00167 

0.000835 

0.1272 

2000 

44467 

44281 

0.00186 

0 . 00093 

0.1416 

84.0 

2400 

44452 

44231 

0.00221 

0.001105 

0.1686 

2800 

44469 

44229 

0.00240 

0.00120 

0.1830 

85-5 

Room  temperature,  76°  F. 
Speed,  325  r.  p.  m. 
Bearing,  No.  308.     Bore  1.5748  in. 
Rated  radial  load,  1450  lb. 


Table  40. — Log  or  Test  on  a  No.  308  Radial  Bearing  under  a  Variable  Radial  Load. 


CONSTRUCTION  OF  BALL  BEARINGS 


I8S 


Room  temperature,  73°.  F. 
Speed   325  r.  p.  m. 
Bearing,  No.  308. 

Table  41.— Log  of  Test  of  a  No.  308  Radial  Bearing  under  a  Constant  Radial  and 

Variable  Thrust  Load. 


Thrust 

Balance 
Weight 

Coefficient  of 

Oil  temperature 

load 
lb. 

Weight 

Average 

Torque 

friction  in  per 
cent. 

degrees  Fahren- 
heit 

2  CO 

0.132 

0.125 

0.128 

1 .262 

0.803 

73 

400 

0.158 

0.14S 

0153 

1 .  509 

0.479 

600 

0.218 

0.201 

0.209 

2.061 

0.436 

76 

800 

0.288 

0.267 

0.277 

2.732 

0.434 

1000 

0-355 

0-343 

0.340 

.      3.440 

0.437 

80 

1200 

0.412 

0.413 

0.412 

4.061 

0.431 

1350 

0.446 

0.485 

0.466 

4.600 

0-433 

85        * 

Room  temperature,  74°  F. 
Speed,  325  p.  m. 
Bearing,  No.  308. 

Table  42. — Log  of  Test  of  a  No.  308  Radial  Bearing  under  a  Variable  Thrust  Load. 

Certain  forms  have  also  been  devised  carrying  two  rows  of  balls,  so  arranged 
that  original  radial  load  develops  thrust  components  and  vice  versa.  Such  a 
bearing  is  thus  capable  of  carrying  also  simultaneous  thrust  and  radial  loads. 
While  ball  bearings  all  demand  a  very  high  degree  of  accuracy,  this  need  is 
greatly  accentuated  for  such  two  row  construction.  A  location  which  demands 
a  bearing  of  this  kind  is  in  the  hubs  of  automobile  wheels. 

A  word  should  be  said  about  the  mounting  of  radial  bearings.     Good  prac- 


i86 


BEARINGS  AND  THEIR  LUBRICATION 


tice  calls  for  a  fit  of  the  inner  race  upon  the  shaft,  such  that  it  will  not  slide  but 
rotate  with  it  as  a  single  member.  Keying  is  objectionable  because  of  the 
weakening  of  the  race  cross-section.  Approved  practice  consists  in  making 
the  inner  race  a  pressing  fit  on  the  shaft  and  binding  it  laterally  by  means  of 
jam  nuts  or  some  equi.valent  device.  A  hammer  should  never  be  directly 
employed  in  contact  with  a  ball  bearing.     Nor  should  the  inner  race  ever  be 


FIG.    73.      HESS-B RIGHT   DOUBLE-ROW  BALL  BEARING. 

forced  home  by  pressure,  or  blows  applied  to  the  outer  one,  acting  through  the 
balls.  If  the  outer  race  slides  within  its  housing  no  harm  is  done.  For  that 
reason  such  a  fit  is  usually  made  a  light  pressing  fit  or  a  "sucking  fit"  with  a 
little  freedom  sidewise.  In  fact  such  sliding  is  beneficial  and  is  generally 
insisted  in  by  ball  bearing  makers  as  the  slow  creep  gradually  brings  all  portions 
of  this  outer  race  around  under  the  load  and  so  distributes  its  effect  over  the 
whole  of  the  outer  race  track. 


CONSTRUCTION  OF  BALL  BEARINGS 
TWO  ROW  RADIAL  BALL  BEARINGS 


187 


For  reasons  cited  on  page  185  several  developments  of  double  row  or 
two-row  bearings  have  appeared.  Fig.  73  shows  such  a  bearing  by  Hess- 
B  right.  This  consists  of  an  inner  race,  an  outer  race,  two  rows  of  balls  and  a 
ball  separator.  The  inner  ball  track  is  concaved,  the  outer  one  flat,  but  inclined 
at  an  angle  of  30°  to  the  axis.  The  inner  race  radius  center  lies  in  the  line 
passing  through  the  ball  center  normally  to  the  outer  race  face.  The  flat  inner 
race  allows  an  even  distribution  of  load  between  the  two  rows. 

Fig.  74  shows  a  bearing  of  this  kind  made  by  the  New  Departure  Manu- 
facturing Company.  The  balls  are  staggered,  fit  into  a  two-piece  outer  race, 
one-piece  inner  race,  and  the  parts  are  assembled  by  a  soft  outer  steel  shell. 


FIG.    74.      TWO-ROW  BALL  BEARING  OF  NEW  DEPARTURE  MANUFACTURING  COMPANY. 


Fig.  75  shows  a  Swedish  bearing  of  the  S.  K.  F.  Ball  Bearing  Company. 
The  novel  feature  is  the  spherical  bearing  surface  of  the  outer  race  which  makes 
the  bearing  self  alining,  for  the  center  of  this  spherical  surface  lies  on  the  axis 
of  the  bearing.  The  method  of  caging  the  balls  is  shown  as  well  as  the  method 
of  assembling. 

It  must  not  be  assumed  that  carrying  capacity  is  doubled  by  doubling  the 
number  of  balls.  Workmanship  of  the  necessary  accuracy  is  not  feasible. 
When  accuracy  of  workmanship  is  depended  upon  solely  for  load  distribution,  a 
liberal  factor  must  be  employed  in  reduction;  when  in  addition  races  of  small  cur- 
vature as  compared  with  the  ball  grooves  are  used  this  factor  must  be  still  further 
liberalized.     See  Section  I,  page  176,  relating  to  curvature  of  race  grooves. 


i88 


BEARINGS  AND  THEIR  LUBRICATION 


CONE-TYPE  BALL  BEARINGS 

Another  type  that  is  used  in  small  sizes,  as  for  magnetos,  is  a  modification  of 
the  cone-type  bearing  of  bicycle  days.  It  is  particularly  serviceable  where 
assembling  is  difficult,  for  the  inner  race  can  be  slipped  into  its  seat,  the  ball 
filled  cage  then  inserted  and  the  outer  race  added. 

This  type  must  be  adjusted  with  care,  but  it  is  a  fallacy  to  think  that  adjust- 
ment can  compensate  for  wear. 

CAGES  FOR  THRUST  BEARINGS 

Experiment  has  shown  that  the  balls  in  thrust  bearings  will  run  against  each 
other  in  a  manner  similar  to  radial  bearings  if  they  are  free,  thus  it  is  customary 
to  cage  the  balls.  .  Mechanically  such  a  cage  has  another  use  in  that  it  assembles 
the  balls  into  a  unit,  permitting  them  to  be  easily  handled  in  the  operations  of 
shipping  and  inserting  in  the  bearing  housing.  In  order  that  each  ball  will 
take  its  proportionate  share  of  the  load,  it  is  essential  that  at  least  one  of  the 
races  should  be  provided  with  a  spherical  groove. 


FIG.    75.       S.    K.    F.      TWO-ROW    RADIAL   BALL    BEARING. 

That  the  bearing  may  adjust  itself  to  its  load,  it  is  common  practice  to  make 
the  lower  race  or  ring  with  a  spherical  surface  entering  a  spherical  seat  in  the 
machine  housing,  or  in  an  adaptor  ring. 

The  advantage  of  the  latter  construction  lies  in  the  difficulty  with  which  a 
spherical  seat  is  machined  in  large  castings. 

Experiments  have  shown  that  the  theoretical  advantages  of  such  a  con- 
struction are  realized  in  practice. 


CONSTRUCTION  OF  BALL  BEARINGS 


189 


FRICTION  OF  THRUST  BALL  BEARINGS 

Fig.  76  shows  a  typical  curve  for  a  thrust  ball  bearing  under  a  variable 
thrust  load. 

Figs.  77  and  78  show  a  two-row  thrust  bearing  made  by  the  S.  K.  F.Bali 
Bearing  Company.  Because  o|  the  spherical  casing  it  is  self  alining.  It  will 
take  thrust  in  either  direction. 

ACCURACY  OF  BALL  BEARINGS 

There  is  no  other  machine  element  produced  in  large  numbers  which  calls 
for  such  accurate  workmanship  as  do  high-grade  ball  bearings.  The  accuracy 
is  almost  painful.     No  details  can  be  overlooked  or  slighted. 

At  the  outset  the  material  must  be  suitable  in  both  balls  and  races,  the  parts 
must  be  properly  hardened  and  tempered,  and  the  grinding  of  the  race  grooves 
and  balls  must  be  accurate  to  within  o.oooi  in.     It  is  a  practical  impossibility 


2^ 


•2  .e  0.11 


a 


0.10 
o.c 

0.08 
0.07 
0.06 
0.05 


" 

~ 

~ 

~ 

~ 

"" 

~ 

~ 

- 

\ 

\ 

s 

V 

^a 

nab 

Jp 

rr 

rhl, 

st 

ji 

21 

iJ 

L-J 

' 

— 1 

p- 

- 

J 

200 


400 


FIG.    76.      TYPICAL   CURVES  /OR   A 


600        800       1000      1200      1400      1600 
Speed  in  R. P.M. 

SMALL    THRUST    BEARING   UNDER    VARLABLE    LOAD. 


commercially  to  grind  all  balls  to  a  limit  of  o.oooi  in.  In  practice,  therefore, 
they  are  sorted  into  grades  and  the  allowable  limit  of  variation  in  each  grade  is 
O.OOOI  in.  When  a  bearing  is  assembled,  care  is  taken  not  to  mix  balls  of 
different  grades.  Only  by  adopting  this  course  is  it  possible  to  produce  a  bear- 
ing in  which  the  load  will  be  properly  distributed  between  the  balls  with  an 
assurance  that  one  ball  if  large  may  not  at  times  take  all  of  the  load,  or,  if  small, 
receive  none,  allowing  its  neighbors  to  carry  its  share.  Furthermore,  it  is  the 
only  way  in  which  to  produce  a  ball  bearing  that  will  run  quietly.  Irregularity 
among  the  balls  themselves  is  a  fruitful  source  of  noise;  possibly  second  only 
to  inaccuracies  in  the  races.  As  a  general  rule,  however,  the  greater  the 
inaccuracies  in  both  races  and  balls,  the  greater  will  be  the  noise  in  running. 
Only  by  having  balls  of  practically  equal  diameter  can  a  bearing  be  made  that 
will  run  quietly. 


IQO 


BEARINGS  AND  THEIR  LUBRICATION 


This  indicates  the  degree  of  accuracy  required  of  the  bearing  manufacturer. 
There  is  another  side  to  the  question,  however,  and  that  is  the  accuracy  of  the 
parts  of  the  machine  in  which  the  bearing  is  fitted. 

If  the  two-part  housing  in  which  the  outer  race  is  held  does  not  fit,  it  is  more 
than  likely  that  if  tightly  clamped,  it  distorts  the  race  and  presses  upon  the 


FIG.  77.   S.  K.  F.  DOUBLE  THRUST  BALL  BEARING. 

balls,  to  the  detriment  of  the  running  of  the  bearing  and  perhaps  causing  its 
failure.  Again,  if  the  shaft  is  too  much  larger  in  diameter  than  the  bore  of  the 
inner  race,  the  latter  may  be  expanded  slightly  and  the  balls  pinched.  These 
are  two  common  causes  of  failures  in  ball  bearings. 

Another  trouble  is  overloading.     This  is  the  chief  cause  of  abrasions.     If 
the  abrasions  are  in  the  middle  of  the  groove  of  either  a  radial  or  thrust  bearing, 


/%JI 


FIG.    78.      SECTION   OF   S.    K.    F.    DOUBLE    THRUST   BALL   BEARING. 

they  probably  indicate  that  the  load  in  a  normal  direction  is  too  great.  If  abra- 
sions are  on  one  side  of  the  groove  in  a  radial  bearing,  they  indicate  that  the 
thrust  load  is  too  great.  If  the  track  of  the  balls  in  a  radial  bearing  swerves 
from  side  to  side,  it  indicates  deflection  of  shaft  or  bearing,  or  original  mis- 
alinement. 


CONSTRUCTION  OF  BALL  BEARINGS  191 

MATERIALS  FOR  BALL  BEARINGS 

The  material  used  for  ball  bearings  for  machinery  is  steel.  In  the  higher 
duty  bearings,  both  balls  and  races  are  made  from  special  alloy  (tool)  steel  that 
has  a  very  high  elastic  limit  after  hardening  and  drawing. 

In  some  bearings,  both  balls  and  races  are  made  of  a  low  carbon  steel 
with  case-hardened  surfaces.  For  severe  service  such  bearings  have  not  proved 
successful.  The  elastic  limit  of  the  material  is  much  lower  than  of  the  alloy 
(tool)  steel,'  and  in  the  ball  there  is  a  tendency  for  the  hard  shell  to  separate 
from  the  softer  core;  this  is  in  evidence  also  in  the  races. 

In  very  large  thrust  ball  bearings  a  high  carbon  unannealed  tool  steel  is 
sometimes  used  without  hardening.  This  is  because  of  the  difficulty  and  expense 
of  hardening  and  grinding  such  large  rings  without  the  accompanying  losses 
from  warping  and  checking  and  the  difficulty  of  getting  perfect  uniformity  of 
surface  hardness.  Hess-Bright  (D.  W.  F.),  however,  produce  races  up  to  54  in. 
diameter  of  a  special  alloy  steel  that  are  hardened  throughout,  as  are  their 
smaller  ones. 

PROPORTIONS  OF  BALL  BEARINGS 

The  proportions  of  ball  bearings  are  well  standardized.  Oddly  enough, 
ball  diameters  are  universally  expressed  in  inches  and  fractions  thereof,  while 
the  dimensions  of  the  races  are  given  in  either  millimeters  or  English  units. 
German  builders  use  millimeters  with  a  single  excej^tion  where  a  firm  has 
developed  a  series  in  English  units,  adapting  them  for  the  British  trade,  though 
that  also  uses  chiefly  ball  bearings  in  millimeters.  Even  in  England  most 
ball  bearings  are  made  to  millimeters  as  is  also  the  more  general  practice  of 
American  manufacturers  who  have  followed  the  German  example.  This  gen- 
eral adoption  of  the  millimeter  dimensions  is  due  to  the  fact  that  early  German 
makers  rehabilitated  the  ball  bearing  by  the  development  of  the  principles  and 
construction  data  of  the  modern  type  and  secured  a  wide  vogue  for  their  pro- 
ducts that  made  the  sizes  standard.  As  a  rule  each  manufacturer  makes  a 
wide  and  narrow  type  of  radial  bearing,  with  three  series  for  each;  namely, 
light,  medium,  and  heavy.  In  some  cases  a  fourth  series  has  been  standardized, 
known  as  '*  extra  heavy. " 

Thrust  bearings  are  usually  made  in  three  series,  light,  medium,  and  heavy. 
The  heavier  the  bearing  the  larger  the  balls  for  a  given  strength. 

At  the  191 1  Spring  Meeting  of  the  Society  of  Automobile  Engineers,  the 
standards  committee  rendered  a  report  containing  Tables  43,  44,  and  45,  giving 
recommended  standard  dimensions  for  the  light,  medium,  and  heavy  series, 
respectively,  for  radial  ball  bearings.  These  standards  are  referred  to  as 
**Ball  Bearings  Standards  A."     It  will  be  noted  that  the  last  column  gives  a 


192 


BEARINGS  AND  THEIR  LUBRICATION 


radial  load  rating  in  pounds  for  each  bearing.  In  this  connection  the  committee 
reported: 

''Attention  is  called  to  the  fact  that  the  capacities  given  in  the  tables  are 
based  upon  ball  bearings  manufactured  of  suitable  workmanship  and  of  suitable 
material  and  running  at  uniform  speed  and  uniform  radial  load,  the  speed  not 
exceeding  500  revolutions  per  minute." 

"  It  is  further  suggested  in  explanation  of  the  load  standards  that  it  cannot 
be  expected  that  all  conditions  will  be  covered  by  the  loads  given.  ^  For  con- 
ditions of  shock,  actual  thrust,  and  a  combination  of  the  two,  greater  factors  of 
safety  will  have  to  be  used." 

Table  46  is  taken  from  this  same  report  and  shows  the  tolerances  or  limits 
worked  to  by  many  bearing  manufacturers  as  embodied  in  statements  made 
to  the  committee.     The  table  applies  to  outer  race  diameter,  bore  and  width. 

Other  standard  dimensions  are  "not  given  here  for  they  can  be  found  in 
manufacturers'  catalogues. 


Table  43. — rEoroRTiONs  of  Radial  Ball  Bearings — Light  Series 


CONSTRUCTION  OF  BALL  BEARINGS 


193 


No.  of 
bearing 

Bore 

Diameter 

Width 

Comer  at  bore  of 
inner  race 

Radial 
load     in 

lbs. 

Mm. 

In. 

Mm. 

In. 

Mm. 

In. 

Mm. 

In. 

300 

10 

0.39370 

35 

1.3779S 

II 

0.43307 

0.04 

200 

301 

12 

0.47244 

37 

1.45669 

12 

0.47244 

0.04 

240 

302 

15 

0.5905s 

42 

1.65355 

13 

0.51181 

0.04 

280 

303 

17 

0.66929 

47 

I .85040 

14 

0.55118 

0.04 

370 

304 

20 

0.78740 

52 

2.04725 

15 

0.59055 

0.04 

440 

305 

25 

0.98425 

62 

2.4409s 

17 

0.66929 

0.04 

620 

306 

30 

1 .18110 

72 

2.83465 

19 

0.74803 

3 

0.08 

860 

307 

35 

I. 37795 

80 

3.14962 

21 

0.82677 

2 

0.08 

IIOO 

308 

40 

1.57481 

90 

3.54332 

23 

0.90551 

2 

0.08 

1450 

309 

45 

I. 77166 

100 

3.93702 

25 

0.9842s 

2 

0.08 

1750 

310 

SO 

I. 96851 

no 

4.33072 

27 

I .06299 

2 

0.08 

2100 

3" 

55 

2. 16536 

120 

4.72443 

29 

1.14173 

2 

0.08 

2400 

312 

60 

2.36221 

130 

5.11813 

31 

1.22047 

2 

0.08 

2800 

313 

65 

2.55906 

140 

5.51183 

33 

I .29921 

3 

0. 12 

3300 

314 

70 

2.75591 

150 

5.90554 

35 

1.37795 

3 

0. 12 

4000 

315 

75 

2.95277 

160 

6.29924 

37 

1.45669 

3 

0. 12 

4400 

316 

80 

3.14962 

170 

6.69294 

39 

1.53544 

3 

0.12 

5000 

317 

85 

3.34647 

180 

7.08664 

41 

I .61418 

3 

0. 12 

5  700 

3i8 

90 

3.54332 

190 

7.48035 

43 

I .69292 

3 

0. 12 

6400 

319 

95 

3.74017 

2PO 

7.87405 

45 

I .77166 

3 

0. 12 

7000 

320 

100 

3.93702 

215 

8.46460 

47 

1.85040 

3 

0.12 

7700 

321 

los 

4.13387 

225 

8.85830 

49 

I. 92914 

3 

0. 12 

8400 

322 

no 

4.33072 

240 

9.44886 

SO 

I. 96851 

3 

0. 12 

1 0000 

Table  44. — Proportions  of  Radial  Ball  Bearings — Medium  Series 


Bore 


Diameter 


Width 


No.  of 
bearing 


Comer  at  bore  of 
inner  race 


Mm. 


In. 


Mm. 


In. 


Mm. 


In. 


Mm. 


In. 


Radial 

load     in 

lbs. 


40, 

17 

0.66929 

62 

2.44095 

17 

0.66929 

I 

0.04 

850 

404 

20 

0.78740 

72 

2.83465 

19 

0.74803 

2 

0.08 

1050 

405 

25 

0.98425 

80 

3.14962 

21 

0.82677 

2 

0.08 

1320 

406 

30 

I . 18110 

90 

3-54332 

23 

0.90551 

2 

0.08 

1600 

407 

35 

1.37799 

100 

3-93702 

25 

0.98425 

2 

0.08 

1900 

408 

40 

I. 57481 

no 

4-33072 

27 

I .06299 

2 

0.08 

2200 

409 

45 

I .77166 

120 

4.72443 

29 

1.14173 

2 

0.08 

2500 

410 

so 

1. 9685 1 

130 

5.11813 

31 

1.22047 

3 

0.08 

3400 

411 

55 

2. 16536 

140 

5.S1183 

33 

I . 29921 

3 

0. 12 

^900 

412 

60 

2.36221 

ISO 

5-905S4 

35 

1.37795 

3 

0. 12 

4400 

413 

6S 

2.55906 

160 

6.29924 

37 

1.45669 

3 

0.12 

4900 

414 

70 

2.75591 

180 

7.08664 

42 

1.6535s 

3 

0. 12 

6200 

41s 

75 

2.95277 

190 

7.4803s 

45 

I. 77166 

3 

0.  13 

6600 

416 

80 

3.14962 

200 

7.87405 

48 

1.88977 

3 

0.  12 

7300 

417 

85 

3-34647 

210 

8.26775 

52 

2.04725 

3 

0.  13 

8580 

418 

90 

3-54332 

22s 

8.85830 

54 

2.12599 

3 

O.I3 

1 0000 

419 

95 

3-74017 

250 

9.84256 

SS 

2.16536 

3 

0.12 

11880 

420 

100 

3.93702 

26s 

10.43311 

60 

2.36221 

3 

O.I3 

14000 

Table  45. — Proportions  of  Radial  Ball  Bearings — ^Heavy  Series. 


13 


194 


BEARINGS  AND  THEIR  LUBRICATION 


Recommended 

by  users                   Rhineland 
Warner  Gear  Co.          S.  R.  Shepard 
C.  E.  Davis 

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SECTION  III 

Typical  Designs' and  Mountings  for  Ball  Bearings 

In  this  section  are  given  a  number  of  typical  designs  and  mountings  for 
ball  bearings  drawn  from  the  practice  of  a  few  manufacturers,  both  foreign 
and  domestic,  and  applying  to  various  kinds  of  machines.  Obviously  other 
firms  could  have  furnished  similar  data  had  there  been  space  to  include  them. 

The  first  examples  are  taken  from  machine  tools  and  indicate  the  interest 
that  their  builders  have  felt  in  adapting  friction  reducing  bearings  to  their 
product.  Figs.  79  to  82,  inclusive,  are  of  grinder  spindles,  ball  bearing  mounted. 
Fig.  79  is  from  an  article  by  W.  M.  Byorkman,  printed  in  the  American 
Machinist  of  February  9,  191 1,  page  242. 


m^mmmmmm^mmm.     | 


FIG.    79.       MOUNTING    OF   GRINDER   SPINDLE    ON    BALL   BEARINGS   AND    FLOATING    BUSHINGS. 

It  shows  a  grinder  spindle  mounted  on  radial  ball  bearings  with  floating 
bushings.  These  bushings  are  used  to  give  steadiness  to  the  spindle  as  in  work 
where  the  dimension  tolerances  are  small.  They  are  located  close  to  the  ball 
bearings  and  have  inside  and  outside  clearances  of  about  0.0002  in.  They  have 
end  clearances  only  sufficient  to  permit  of  appreciable  movement. 

The  bushings  carry  no  load,  but  the  clearance  spaces  are  filled  with  oil  which 
acts  like  a  dash  pot  to  suppress  the  minute  vibrations  which  would  otherwise 
arise  in  the  ball  bearings.  The  final  effect  is  to  do  away  with  the  "waves"' 
frequently  noticed  on  ground  work  where  the  grinding  spindle  is  ball  bearing 
mounted 

195 


196 


BEARINGS  AND  THEIR  LUBRICATION 


The  endwise  movement  of  the  shaft  is  prevented  by  a  pair  of  thrust  ball 
bearings  seen  at  the  right  hand  end. 

This  particular  spindle  arrangement  v^^as  applied  to  a  grinding  machine  for 
sizing  to  thickness  type  bars  of  a  typewriter. 


Adjusting  Screw 


FIG.    80.       SECTION    OF   SPINDLE    BEARINGS    OF   RIVETT    GRINDER. 

Fig.  80  shows  in  section  the  spindle  and  bearings  of  a  Rivett  grinder.  From 
a  test  made  by  A.  Morin  in  Paris,  France,  a  speed  of  120,480  revolutions  per 
minute  has  been  obtained  with  a  spindle  of  one  of  these  machines.     This  is 


FIG.    81.      BALL-BEARING    MOUNTINGS    FOR  HIGH   SPEED    SPINDLE   AND    ITS    PULLEY. 

undoubtedly  the  highest  speed  ever  recorded  for  a  rotating  machine  element 
driven  by  mechanical  power. 

Figs.  81  and  82  show  two  recommended  grinder  spindle  mountings  developed 

Endclosure  Lips  Bored  Vei  Inch  larger  than  Shaft. 

Inner  Races  of  Radial  Bearings  Light  Drive  Fit  on  Shaft. 

Outer  Races  of  Radial  Bearings  Sucking  Fit  in  Housing. 


Washers  for 

Adjustment  of 

Thrust  Bearings. 

FIG.    82.      BALL-BEARING    MOUNTINGS    FOR   BUFFING    OR   GRINDING   HEAD. 

by  the  Hess-Bright  Manufacturing  Company.  In  Fig.  81  it  will  be  noticed  that 
the  driving  pulley  is  carried  on  two  radial  bearings  supported  by  a  sleeve  inde- 
pendent from  the  spindle  itself,  thus  the  stress  of  the  belt  pull  does  not  reach  the 


TYPICAL  DESIGNS  AND  MOUNTINGS 


197. 


spindle.  Another  feature  is  in  the  right  hand  radial  bearing  of  the  spindle 
itself.  The  outer  race  instead  of  having  a  groove  for  the  balls  has  a  plain  cylin- 
drical surface.  Thus  the  balls  have  a  chance  to  move  in  the  direction  of  the 
length  of  the  spindle  if  the  parts  lengthen  from  heat. 


c — ^sSS^ 


83.      BALL-BEARING    THRUST   MOUNT 
INGS    FOR   DRILLER   SPINDLES. 


BALL-BEARING   MOUNTING    FOR   A 
VERTICAL   SPINDLE. 


Fig.  82  needs  but  little  comment  except  to  point  out  the  grooves  at  the  ends 
of  the  bearing  housings  provided  to  prevent  the  escape  of  oil  and  the  entrance 
of  dust. 


Top  Bearing  if  Pulley 
is  Placed  on  Bottom 


Top  Bearing        .^rrrm 

if  Pulley    ' ^^ 

is  Placed   ,,-- 
on  Top 


Bottom  Bearing  if  Pulley 
is  Placed  ou  Bottom 


Bottom  Bearing  if  Pulley 
is  Placed,  on  Top 


85.      BALL-BEARING   MOUNTING    FOR 
ELECTRIC    MOTOR   SHAFT. 


FIG. 


86.      BALL-BEARING   MOUNTINGS   FOR 
VERTICAL   ELECTRIC    MOTOR. 


Fig.  83  shows  two  mountings  for  driller  spindles  and  is  likewise  recommended 
by  the  Hess-Bright  Manufacturing  Company. 

Fig.  84  shows  a  vertical  spindle  resting  on  a  ball  thrust  at  the  bottom  and 
guided  by  two  radial  bearings.     It  is  from  designs  of  the  Rhineland  Machine 


198 


BEARINGS  AND  THEIR  LUBRICATION 


FIG.    87.      BALL-BEARING   MOUNTING    FOR   STREET 
RAILWAY   MOTOR   SHAFT. 


FIG.  88.      BALL-BEARING  MOUNT- 
INGS   FOR   TRUCK   WHEEL. 


FIG.    89.      BALL-BEARING    MOUNTING   FOR  FIG.    90.       CENTRIFUGAL  OILING  DEVICE  AND 

WORM   AND    WORM   WHEEL.  BALL-BEARING  MOUNTING  FOR  A  SHAFT  STEP. 


FIG.    91.      BALL-BEARING    SHAFT    MOUNTING   WITH   SELF-ALINING    SEAT. 


TYPICAL  DESIGNS  AND  MOUNTINGS 


199 


Works  Company.  An  important  point  in  installing  thrust  bearings  is  this:  If 
the  upper  race  is  tight  on  the  shaft,  the  lower  leveling  washer  must  be  free  to  a 
slight  extent  to  allow  for  movement  in  preserving  alinement.  On  the  other 
hand,  if  the  upper  washer  is  made  with  a  slight  clearance  over  the  shaft  diameter 
the  lower  ring  need  not  be  free. 


FIG.    92.      BALL-BEARING   PILLOW   BLOCK. 


FIG.    93.      TWO    FORMS    OF    SHAFT 
ADAPTERS  FOR  BALL  BEARINGS. 


MOTOR  BALL  BEARINGS 

Figs.  85,  86  and  87  show  three  applications  of  ball  bearings  to  the  shafts 
of  electric  motors.  Fig.  85  is  a  Ivhineland  design  for  a  small  sized  motor.  The 
clearance  in  an  axially  direction  of  the  bearing  at  the  right  hand  is  very  plainly 


FIG.    94.      BALL-BEARING   SWIVEL   FOR  A   LIFTING   EYE. 

shown.  Fig.  86  shows  two  mountings  for  the  shaft  of  a  45-h.  p.  vertical 
motor  involving  two  different  combinations  of  radial  end  thrust  bearings.  It  is 
from  the  practice  of  the  Standard  Roller  Bearing  Company. 

Mountings  of  ball  bearings  for  street  railway  motors  are  shown  in  Fig.  87. 
A  large  number  of  motors  with  this  type  of  mounting  have  been  in  use  on  some 
of  the  street  railway  lines  in  New  York  City.     The  design  is  Hess-B right. 


200 


BEARINGS  AND  THEIR  LUBRICATION 


FIG.    95.      BALL-BEARING    MARINE   PROPELLER  BLOCK. 


Oil  Returning  Grooves  A 


Screw  for  Locating  Leveling 
Washer  so  that  Grooves  A 
are  kept  Coincident  with 
Sloping  Channels  in  Casing 
Walls.  Required  for  Returning 
of  Oil  to  Center  of  Beariag 


Gouge  Slot  Across  Oil 
Retaining  Grooves  at  Convenient 

,ce  to  Afford  Sharp  Edge 
fur  Wiping  Oil  from  Shaft 


FIG.    96.      COMBINATION    RADIAL   AND    THRUST   BALL-BEARING   MOUNTING. 


FIG.    97.      BALL-BEARING    MOUNTING    FOR   THREE-BEARING,    FOUR-CYLINDER   AUTOMOBILE 

CRANK   SHAFT. 


TYPICAL  DESIGNS  AND  MOUNTINGS  201 

GENERAL  APPLICATIONS 

Figs.  88  to  96,  inclusive,  show  a  number  of  ball  bearing  mountings  applied 
to  machinery  in  general.  Of  these  Figs.  88  to  94,  inclusive,  are  Rhineland 
designs;  Fig.  95  is  a  Hess-Bright  design  and  Fig.  96  a  Standard  Roller  Bearing 
design. 


FIG.    98.    BALL-BEARING    MOUNTING    FOR   TWO-BEARING,    FOUR-CYLINDER  AUTOMOBILE 

CRANK    SHAFT. 

Fig.  88  shows  two  radial  bearings  applied  to  a  heavy  truck  wheel.  Fig.  89 
presents  the  mounting  of  a  worm  and  worm  wheel  with  both  radial  and  thrust 
bearings.  The  thrust  bearings  are  of  the  Rhineland  double  thrust  type.  Fig. 
90  is  of  interest  for  the  design  of  a  centrifugal  oiler  in  a  chamber  below  the  step 
whereby  oil  is  pumped  from  the  lower  level  of  this  chamber  upward  through 


FIG.  99.       BALL-BEARING  AUTOMOBILE  mJB 
MOUNTING — FLOATING   TYPE. 


FIG. 


100.  BALL- BEARING  AUTOMOBILE  HUB 

MOUNTING — FIXED   TYPE. 


the  center  of  the  shaft  and  then  outward  through  channels  into  the  spaces  sur- 
rounding the  balls.  The  oil  finds  its  way  back  to  the  reservoir  through  an 
annular  space  around  the  bearing. 

Figs.  91  and  92  are  pedestal  box  mountings.  In  Fig.  91  the  housing  for  the 
bearings  has  a  spherical  mid-section  fitting  a  spherical  seat  in  the  pedestal 
frame.     This  allows  for  self  alinement.     Fig.  92  shows  the  outer  races  of  the 


202 


BEARINGS  AND  THEIR  LUBRICATION 


ball  bearings  provided  with  spherical  surfaces  resting  against  cylindrical  sur- 
faces inside  of  the  pedestal  frame.  It  will  be  noted  that  the  inner  surface  of 
the  inner  ball  races  is  conical  and  is  fitted  to  the  cylindrical  shaft  by  means  of 
an  adapter  bushing. 

Fig.  93  shows  iwo  adapter  constructions  of  Hess-Bright  design.  The  one 
at  the  right  has  a  combination  of  two  tapered  bushings,  one  surface  of  each 
fitting  the  cylindrical  shaft  and  the  cylindrical  surface  of  the  inside  of  the  inner 
race,  respectively. 


FIG.    lOI.      BALL-BEARING    MOUNTING   FOR  AUTOMOBILE    REAR   AXLE    DIFFERENTIAL  AND   HUB, 

Fig.  94  shows  a  thrust  bearing  mounted  in  the  swivel  of  a  large  lifting  eye. 
The  following  figure,  Fig.  95,  illustrates  a  marine  propeller  box  fitted  with  two 
thrust  bearings  and  two  radial  bearings  to  carry  the  load  of  the  shaft.  Further 
description  is  unnecessary. 

The  design  of  Fig.  96  is  a  large  thrust  bearing  for  a  2  1/2  in.  shaft  combined 
with  two  radial  hearings  to  carry  vertical  load. 


A  FEW  AUTOMOBILE  APPLICATIONS 

Figs.  97  to  1 01  inclusive,  show  a  few  automobile  ball  bearing  mountings. 
Of  these,  .Fig.  97  is  from  the  Rhineland  Company  and  the  others  from  Hess- 
Bright. 

Fig.  97  shows  three  bearings  for  the  crankshaft  of  a  four-cylinder  automobile. 
Each  bearing  consists  of  two  radial  bearings  placed  side  by  side  and  having  the 


TYPICAL  DESIGNS  AND  MOUNTINGS 


203 


external  surface  of  the  outer  races  of  each  pair  ground  to  the  same  radius  of 
curvature.     This  makes  each  pair  self  alining  and  is  a  form  of  two-row  ball 
bearing  not  mentioned  in  the  previous  discussion  of  this  subject. 
The  other  figures  in  this  group  do  not  need  particular  mention. 


FIG. 


102.       BALL  THRUST  BEARING 
WITH   FLAT   WASHERS. 


FIG.  103. 


BALL  JOURNAL  BEARING  FOR 
FLOATING  SHAFT. 


The  final  construction,  Figs.  102  and  103,  are  from  the  Standard  Roller 
Bearing  Company  and  show,  respectively,  a  plain  ball  thrust  having  flat  washers — 
a  type  previously  described — and  a  special  ball  journal  bearing  developed  for 
use  on  a  spindle  which  had  considerable  endwise  travel. 


SECTION  IV 

Lubrication  of  Ball  Bearings 

No  better  discussion  of  the  lubrication  of  ball  bearings  can  be  found  than 
is  given  in  an  article  in  the  American  Machinist,  Vol.  XXXIV,  page  87,  by 
W.  L.  Batt.     It  is  given  herewith. 

Oil  is  essential  for  the  operation  of  plain  bearings,  but  it  is  not  so  well 
known  that  ball  bearings  likewise  require  a  certain  amount  of  lubrication;  it  is 
all  too  frequently  assumed  that  the  latter  may  run  entirely  dry. 

Contrary  to  this  assumption  lubrication  has  a  very  positive  action  in  a  ball 
bearing.  This  action  is  many  fold,  existing  both  between  the  ball  and  the 
path  upon  which  it  rolls  and  between  the  ball  and  the  separator.  Though 
the  principle  of  the  ball  bearing  is  a  rolling  one,  there  is  in  the  most  perfectly 
designed  bearing  a  sliding  action  which,  however  slight,  must  yet  be  provided 
for.  If  balls  and  raceways  were  absolutely  incompressible  there  would  in 
practice  be  solely  point  contact  and,  therefore,  a  pure  rolling  action,  but  as  far 
as  present  day  mechanical  processes  go  even  the  most  perfect  of  materials  are 
somewhat  elastic  and  there  is  an  actual  area  of  deformation  between  the  ball 
and  the  raceway  under  load.  At  the  extremities  of  this  area,  vs^hose  diameter 
is,  of  course,  very  small,  sliding  must  exist,  but  the  presence  of  a  lubricant  at 
this  point  renders  what  small  amount  of  friction  there  is  almost  negligible. 
Improbable  as  it  may  seem,  there  is  an  actual  film  of  lubricant  maintained 
between  these  surfaces  of  contact,  however  minute  they  are,  providing  the 
lubricant  be  one  of  sufficient  viscosity. 

PROTECTIVE  VALUE  OF  THE  LUBRICANT 

Aside  from  its  lubricating  quality,  oil  or  grease  acts  as  a  protecting  agent 
for  the  ball  bearing.  The  finely  polished  surfaces  of  balls  and  raceways -are 
subject  to  attack  by  rust  through  atmospheric  action  and  may  be  damaged  by 
the  entrance  of  foreign  matter  from  the  outside,  such  as  grit  and  dust.  Lubri- 
cant effectually  surrounding  the  bearing  will  not  be  penetrated  by  this  atmos- 
pheric moisture  and  thus  the  surfaces  of  action  are  preserved  in  their  original 
finely  polished  condition.  If,  however,  rust  or  grit  should  be  allowed  to  attack 
the  carrying  surfaces  these  are  quickly  destroyed  and  an  earlier  failure  of  the 
bearing  may  be  expected. 

In  order  that  the  lubricant  may  in  itself  be  no  source  of  danger,  it  must 

204 


LUBRICATION  OF  BALL  BEARINGS  205 

respond  to  certain  requirements;  incidentally  requirements  exacted  of  the  ball- 
bearing lubricant  are  thoroughly  desirable  in  one  to  be  used  in  plain  bearings 
as  well. 

REQUIREMENTS  OF  A  BALL  BEARING  LUBRICANT 

First  and  most  vital  is  the  requirement  that  the  lubricant  of  itself  shall  do  no 
damage  to  the  bearing,  neither  originally  nor  through  deterioration.  The  most 
common  fault  in  oils  is  the  presence  of  free  acid  or  its  development.  Free 
acid  is  never  found  in  properly  refined  mineral  or  hydro-carbon  oils,  nor  will 
these  deteriorate  due  to  the  action  of  the  atmosphere,  and  therefore  the  only 
kind  of  oil  satisfactory  for  ball  bearing  lubrication  is  mineral  oil.  Vegetable 
oils,  such  as  castor,  cotton-seed,  rape,  linseed  and  the  like  are  barred,  principally 
because  of  their  tendency  to  gum  up,  become  rancid  and  develop  acid.  The 
animal  oils  are  objectionable  for  the  same  reason.  A  simple  test  for  the 
presence  of  objectionable  elements  in  an  oil  or  grease  is  to  coat  a  finely  pol- 
ished surface  of  steel  or  brass  or  the  surface  of  a  ball  bearing,  with  the  oil  or  grease 
in  question  and  expose  it  to  the  sunlight  for  a  few  weeks;  at  the  end  of  that  time 
the  steel  should  show  its  original  polish.  Any  discoloration  or  drying  up  of 
the  lubricant  on  the  steel  is  sufficient  to  bar  it  from  consideration.  The  so-called 
petroleum  greases,  among  which  are  vaseline  and  cosmoline,  have  no  detri- 
mental action  in  themselves,  since  they  are  derivatives  of  mineral  oil ;  they  have, 
however,  low  viscosity  and  a  low  melting-point — 100  to  125°  F.;  they  pound 
out  thin  in  action  and  have  little  lubricating  value.  Their  use  is  limited  to  very 
slow  speeds  and  their  chief  advantage  is  low  cost. 

Just  as  acid  is  a  thing  to  be  guarded  against  in  oils,  so  is  tree  alkali  the 
most  common  enemy  to  ball  bearings,  among  the  greases.  The  familiar  yellow 
cup  grease  is  usually  a  combination  of  a  mineral  oil  and  some  vegetable  or 
mineral  oil  or  fat  which  latter  is  saponified  by  the  addition  of  a  caustic;  the 
result  is  a  lubricant  having  body  and  stiffness  The  saponifying  material 
should  be  small  in  quantity  and  very  carefully  compounded,  else  free  alkali, 
having  a  detrimental  action  on  the  steel  may  result  The  action  of  an  alkali, 
as  of  an  acid,  is  to  pit  or  etch  the  surfaces  upon  which  it  is  deposited. 

The  addition  of  mica,  ground  cork,  wood  and  such  substances,  frequently 
added  to  overcome  noise  in  gear  cases  of  automobiles,  is  a  positive  menace  to 
the  ball  bearing,  since  this  foreign  matter  opposes  free  ball  rotation;  if  it  be 
present  in  large  enough  amount,  the  result  may  easily  be  that  the  balls  are 
wedged  between  the  raceways  and  actual  fracture  may  result.  Certainly  the 
free  rolling  quality  of  the  ball  bearing  will  be  lost. 

The  question  of  the  beneficial  effect  of  graphite  in  ball  bearing  lubrication 
is  one  often  asked.  The  answer  is  simply  that  graphite  in  any  shape  or  form 
that  will  settle  and  pack  with  time  when  quiescent  cannot  be  of  assistance  to 


2o6  BEARINGS  AND  THEIR  LUBRICATION 

the  ball  bearing  itself.  Whether  it  is  a  detriment  of  enough  importance  to 
offset  its  undoubted  advantage  to  gears  or  whatever  the  mounting  may  be, 
depends  entirely  on  the  kind  and  .fineness  of  the  graphite  and  the  proportions 
in  which  it  is  added  to  the  lubricant. 

It  may  readily  be  seen  that  any  considerable  amount  of  graphite  settling  in 
the  bottom  of  the  raceway  will  interfere  with  the  free  rolling  of  the  balls  in 
the  same  manner  as  ground  cork  or  any  other  substance.  The  claim  made 
for  graphite,  and  one  certainly  proved  in  the  case  of  the  plain  bearing,  is  that  it 
deposits  itself  in  a  fine  film  on  the  surface  upon  which  the  shaft  may  revolve. 
A  microscopic  examination  of  the  balls  and  grooves  of  a  carefully  made  ball 
bearing  shows  almost  no  scratches  or  inequalities;  after  such  a  bearing  has  been 
run  in  oil  and  graphite  of  proper  proportions,  examination  again  will  show  that 
absolutely  no  graphite  has  been  deposited.  Frequent  observations  carefully 
made  show  a  slight  increase  in  the  friction  when  graphite  is  present  over  that 
when  oil  alone  is  used.  Such  results  clearly  show  that  graphite  has  no  value 
in  so  far  as  reduction  of  ball  bearing  friction  may  be  concerned.  Over  the 
usual  life  of  a  ball  bearing,  however,  probably  no  actual  damage  will  result 
and  considerable  good  may  have  accrued  to  the  elements  carried  upon  the  ball 
bearings;  such  use  would,  of  course,  be  justifiable.  The  damage,  however, 
from  graphites  that  settle  and  accumulate  in  the  bearings  cannot  be  overlooked. 

EFFECT  OF  SPEED  UPON  CHOICE  OF  LUBRICANT 

The  effect  of  speed  upon  the  choice  of  lubricant  must  likewise  be  considered. 
As  a  general  statement,  greases  are  desirable  for  use  up  to  just  as  high  a  speed 
as  is  possible.  An  undue  increase  of  temperature  and  a  thinning  out  of  grease 
will  indicate  the  limit  of  its  usefulness;  when  this  is  noted  oil  must  be  employed. 
Oil  has  a  higher  viscosity  than  grease  and  theoretically  is  a  better  lubricant, 
but  its  advantage  in  this  regard  is  more  than  offset  by  the  difficulty  of  its  proper 
retention,  in  the  bearing  case.  Such  case  may  be  solidly  packed  with  grease 
that  will  thus  not  only  lubricate  the  bearing,  but  will  also  prevent  foreign 
substances  from  entry.  If  oil  be  used  the  ball  will  merely  dip  into  the  oil,  splash- 
ing it  over  the  entire  bearing,  but  the  bearing  is  then  more  open  to  the  entry  of 
foreign  matter.  There  will  be  more  internal  friction  in  the  grease  than  in  the 
oil  and  it  is  this  that  at  high  speed  produces  excessive  temperature.  As  a 
general  thing  greases  may  be  used  up  to  about  2000  to  2500  revolutions  per 
minute. 

MOUNTINGS  MADE  TO  RETAIN  LUBRICANT 

In  order  that  the  lubricant  may  be  effectually  retained,  various  arrange- 
ments are  used,  depending  upon  the  conditions  surrounding  .the  bearing.  In 
the  simplest  sort  of  mounting  for  a  radial  bearing  the  shaft  projects  through 


LUBRICATION  OF  BALL  BEARINGS  207 

the  casing,  and  the  casing  itself  is  provided  with  two  lips,  between  which  is  a 
space  for  lubricant.  If  the  condition  be  such  that  additional  protection  is 
needed,  one  additional  groove  is  provided  and  this  may  be  fitted  with  a  cup  of 
some  sort  from  which  grease  will  be  steadily  fed  to  the  groove  to  keep  that 
filled.  This  makes  a  definite  frictionless  packing.  For  still  more  severe  con- 
ditions a  third  groove  is  added;  in  this  latter  groove  felt  is  occasionally  placed, 
whose  adherence  to  the  shaft  is  guaranteed  by  some  sort  of  spring  tension. 
Unfortunately,  this  is  subject  to  drying  out  and  thus  loses  its  efficiency;  when 
that  occurs  it  is  a  positive  detriment. 

The  single  or  multiple  groove  arrangement  empty  or  with  only  grease 
filling  is  the  most  effective;  but  it  is  essential  that  the  bore  of  the  lips  be  not 
more  than  1/64  in.  larger  than  the  shaft  in  diameter,  that  the  lip  edges  be  sharp 
instead  of  rounded  over,  and  that  the  lips  be  at  least  3/32  in.  wide.  The  grooves 
may  also  be  cut  in  the  shaft,  leaving  bands  between  with  sharp  edges.  The  only 
objection  to  this  is  the  weakness  of  the  shaft. 

HINTS  ON  THE  CARE  OF  AUTOMOBILE  BALL  BEARINGS 

The  preceding  paragraphs  point  out  the  need  of  lubrication  and  the  necessity 
of  keeping  ball  bearings  free  from  all  foreign  particles.  V.  W  Page,  writing 
in  the  American  Machinist ,  gives  a  number  of  hints  on  the  care  of  ball  bearings 
in  automobiles.  He  says:  "Many  cases  of  trouble  in  rear  axle  and  transmission 
case  bearings  have  been  traced  to  the  presence  of  minute  particles  of  metal 
ground  off  from  the  gears,  or  to  particles  of  sand  loosened  from  the  interior  of 
the  gear  case  or  rear  axle  housing  castings.  In  sliding  gears,  especially  when 
operated  by  an  inexperienced  person,  there  is  a  constant  clashing  of  the  pinions 
in  changing  speeds,  which  tends  to  loosen  particles  of  metal  from  the  teeth. 
These  fall  into  the  lubricants  and  ofttimes  find  their  way  into  the  ball  races. 
The  rapid  failure  of  the  bearing  is  the  inevitable  consequence.  Likewise  failures 
in  ball  bearings  on  automobile  engine  crankshafts  have  been  traced  to  the  pres- 
ence of  particles  of  carbon  and  other  foreign  matter  in  the  oil. 

Another  source  of  trouble  is  the  washing  of  rust  and  dirt  into  the  bearings 
through  carelessness  in  washing  the  automobile.  This  is  particularly  true  of 
wheel  bearings,  where  many  failures  have  been  traced  directly  to  the  presence 
of  rust  and  dirt  caused  by  the  indiscriminate  use  when  washing  with  a  stream  of 
water  under  pressure  of  40  to  50  lb.  per  square  inch;  that  is,  the  normal  pressure 
of  the  city  water  mains. 

Another  practice  that  is  detrimental  to  ball  bearings  is  used  in  many  motor 
car  repair  shops.  It  consists  in  dipping  the  bearings  for  the  purpose  of  cleansing 
into  dirty  gasoline  in  which  perhaps  gears  and  other  parts  covered  with  metallic 
particles  have  already  been  washed. 


2o8  BEARINGS  AND  THEIR  LUBRICATION 

Still  another  practice  prevalent  not  only  in  garages  and  repair  shops,  but  in 
too  many  machine  shops  is  the  employment  of  a  hammer,  particularly  a  hard 
hammer,  in  assembling  or  dismounting  ball  bearings.  It  will  be  remembered 
that  these  are  made  of  hardened  steel.  The  man  who  hammers  a  reamer  or 
cutter  is  immediately  discharged,  but  gets  only  a  reprimand  when  he  similarly 
abases  a  ball  bearing. 


SECTION  V 

Roller  Bearings  with  Flexible  Rollers 

Roller  bearings  with  hollow,  cylindrical,  helical,  flexible  rollers  made  by  the 
Hyatt  Roller  Bearing  Company  are  used  in  machinery  in  general,  on  line 
shafting  and  in  automobiles.  The  rollers  are  wound  from  flat  strip  steel  into 
a  closed  helical  coil.  Thus  they  are  flexible  and  can  adapt  themselves  to  slight 
irregularities  in  either  journal  or  box  without  causing  excessive  pressure.  The 
cylindrical  hollows  in  rollers  serve  as  storage  space  for  lubricant  and  the  helical 


FIG.    104. 


FLEXIBLE    ROLLER    BEARING    MOUNTINGS    FOR   TRANSMISSION   SHAFTS    OF 
MULTIPLE    DRILLER. 


interstices  distribute  the  oil  the  entire  length  of  the  box.    One-half  of  the  rollers 
in  a  box  have  a  right-hand  helix,  the  other  half  left-hand. 

In  the  form  of  bushings,  two  types  are  made,  known  as  the  standard  type 
and  the  high  duty  type.     In  the  standard  type  the  rollers  are  of  carbon  steel  with 
14  209 


2IO 


BEARINGS  AND  THEIR  LUBRICATION 


an  outer  shell  or  lining  of  special  analysis  sheet  steel.  The  rollers  run  in  con- 
tact with  the  shaft  or  journal.  Where  the  bearing  surface  is  generous  it  is 
satisfactory  to  operate  the  rollers  direct  on  soft  steel  surfaces. 

In  the  second  type  the  rollers  are  of  nickel  steel,  31/2  per  cent,  nickel,  and 


FIG.    105.      FLEXIBLE   ROLLER  BEARINGS   MOUNTING   FOR  A   CIRCULAR  SAW  ARBOR. 

heat  treated.  The  lining  is  tubular  and  a  tubular  sleeve  is  provided  to  slip 
over  the  shaft  or  journal.  Both  of  these  parts  are  also  heat  treated.  Several 
devices  are  used  to  cage  the  rollers,  which  are  squared  on  the  ends  to  thrust 
against  the  ends  of  the  box.  As  the  allowable  unit  bearing  pressures  of  the 
high  duty  rollers  are  higher  than  for  the  standard  rollers  the  high-duty  bearings 


m. 


^W 


FIG.    106.       FLEXIBLE   ROLLER  BEARING  APPLIED   TO   COUNTERSHAFT   BEARING  AND   LOOSE 

PULLEY. 


have  the  practical  advantage  of  being  shorter  for  the  same  load  than  the  com- 
mercial bearings. 

The  shafting  bearing  is  of  the  standard  type  with  a  horizontally  split  box 
so  that  it  can  be  put  on  a  shaft  anywhere  and  does  not  have  to  be  slipped  on 
over  the  end. 


ROLLEI^  BEARINGS  WITH  FLEXIBLE  ROLLERS 


211 


FIG.    107.       FLEXIBLE    ROLLER   BEARING    MOUNTING    FOR   CLUTCH   PULLEY. 


FIG.    108.      FLEXIBLE    ROLLER   BEARING    FOR   FLOUR   MILL. 


212 


BEARINGS  AND  THEIR  LUBRICATION 


COEFFICIENTS  OF  FRICTION 

In  the  Trans.  A.S.M.E.,  Vol.  XXVII,  page  504, Prof.  A.  L  Williston  gives 
the  results  of  various  experiments  with  Hyatt  bearings.  From  that  source, 
the  following  coefficients  of  friction  are  taken:  Average  for  various  loads  at 
130  r.  p.  m.,  0.011-4;  at  302  r.  p.  m.,  0.0099;  ^t  585  r.  p.  m.,  0.0147.  The 
average  for  very  heavy  loads  at  185  r.  p.  m.  was  o.oii;  at  215  r.  p.  m.  0.0093. 

A  test  made  by  the  author  comparing  Hyatt  standard  lineshaft  bearings  with 
babbitted  bearings  gave  a  saving  for  the  roller  bearings  of  64.9  per  cent.  The 
shaft  was  152  ft.  long,  2  15/16  in.  in  diameter  and  was  supported  by  20  hangers. 


ALLOWABLE  PRESSURES  AND  SPEEDS 

The  allowable  pressures  for  the  standard  or  commercial  type  are  fixed  by 
the  quality  of  the  shaft  or  journal  against  which  the  rollers  bear.     For  low  speeds 


riG.    log.      FLEXIBLE    ROLLER   BEARING   FOR  PIG   BREAKER. 

up  to  50  r.  p.  m.  the  maximum  limit  lies  between  400  to  500  lb.  per  square 
inch  of  projected  journal  area.  The  method  of  housing,  particularly  in  its 
relation  to  the  distribution  of  load,  and  quality  of  lubrication  have  an  in- 
fluence in  determining  these  limits. 

With  an  increase  of  speed  the  allowable  pressure  decreases.  For  lineshaft 
bearings  up  to  3  15/16  in.  in  diameter  running  up  to  600  r.  p.  m.  30  lb.  per 
square  inch  is  considered  good  practice.  For  larger  shafts  the  same  factor  is 
allowable  up  to  speeds  of  400  r.  p.  m. 

The  high-duty  bearings  carry  much  greater  unit  loads.  A  rating  of  750  lb. 
per  square  inch  at  1000  r.  p.  m.  is  conservative. 

A  limiting  maximum  speed  for  the  standard  or  commercial  bearings  is 
about  1500  r.  p.  m.;  for  the  high-duty  bearings,  3000  r.  p.  m. 


ROLLER  BEARINGS  WITH  FLEXIBLE  ROLLERS 


213 


LUBRICATION 

These  bearings  must  be  well  lubricated.  If  the  speed  is  medium  or  high, 
a  good  body  machinery  oil  is  suitable.  If  the  speed  is  slow  a  heavy  body 
machinery  oil  or  grease  is  better. 


FIG.    no.      FLEXIBLE   ROLLER  BEARINGS   MOUNTING   FOR   SHAFTS   OF  A   VERTICAL   PUMP. 

TYPICAL  APPLICATIONS  IN  MACHINERY 

Several  typical  designs  and  machinery  applications  are  shown  in  Figs.  104 
to  no,  inclusive.  The  first,  Fig.  104,  is  of  shaft  mountings,  for  a  multiple- 
spindle  driller;  the  second,  Fig.  105,  is  a  circular  saw  arbor  on  two  bearings. 


214 


BEARINGS  AND  THEIR  LUBRICATION 


FIG.  III.   STANDARD  FLEXIBLE  ROLLER  BEARING  APPLIED  TO  BEVEL  GEAR  REAR  AXLE 
WITH  BEARING  ON  DIFFERENTIAL  HUB. 


FIG.  112.   STANDARD  FLEXIBLE  ROLLER  BEARING  APPLIED  TO  BEVEL  GEAR  REAR  AXLE  WITH 
BEARING  MOUNTED  ON  SOFT  AXLE. 


FIG.  113.   HIGH  DUTY  FLEXIBLE  ROLLER  BEARING  MOUNTING  ON  BEVEL  GEAR  REAR  AXLE 

WITH  BEARING  ON  TUBE. 


ROLLER  BEARINGS  WITH  FLEXIBLE   ROLLERS 


215 


riG.    114.      HIGH   DUTY   FLEXIBLE    ROLLER  BEARING   MOUNTING   FOR   BEVEL   GEAR   REAR  AXLE 
WITH    SLEEVE   AND    NO    THRUST   BEARING, 


^>//////^//////////////, 


Y/////////////////////A 


FIG.    115.      HIGH    DUTY    FLEXIBLE    ROLLER    BEARING    MOUNTING    FOR    INNER    END    OF    BEVEL 
PINION    SHAFT — MATING    GEAR   ON    REAR  AXLE. 


FIG.    116.      HIGH-DUTY   FLEXIBLE   ROLLER   BEARINGS   IN   SELECTIVE   TRANSMISSION   MOUNTED 

INDEPENDENTLY. 


2l6 


BEARINGS  AND  THEIR  LUBRICATION 


Figs.  io6  and  107  show  a  countershaft  bearing  and  loose  pulley  mounting, 
and  two  bearing  applied  to  a  friction  clutch.  In  this  latter  example  the  short 
high-duty  bearing  carries  the  load  and  the  longer  standard  bearing  is  a  steady- 
ing bearing  at  the  end  of  the  hub. 

Fig.  108  is  a  flour  mill  construction  and  is  of  interest  as  showing  the  device 
used  to  prevent  the  escape  of  oil. 

Fig.  109  is  the  design  of  an  8-in.  bearing  for  a  pig  breaker.  Because 
of  the  length  — 22  in. — two  sets  of  rollers  are  used.  Fig.  no  shows  the  mount- 
ing of  an  8-in.  vertical  spiral  pump.  Roller  bearings  are  used  for  the 
guide  bearings  and  ball  bearings  for  the  thrusts. 


FIG.    117.      HIGH-DUTY    FLEXIBLE    ROLLER    BEARINGS    IN    SELECTIVE    TRANSMISSION    ON 

REAR  AXLE. 


IN  AUTOMOBILE  PRACTICE 

These  bearings  have  had  an  extensive  use  in  automobiles.  The  seven 
engravings,  Figs.  1 1 1  to  117  inclusive,  show  a  few  typical  arrangements.  Figs. 
Ill  and  112  are  of  standard  bearings  applied  to  bevel  gear  rear  axles.  Figs. 
113  and  114  are  of  high  duty  bearings  also  applied  to  bevel  gear  rear  axles. 
Fig.  115  shows  the  inner  bearing  of  a  bevel  pinion  with  a  ball  thrust,  and  Figs. 
116  and  117  show  two  mountings  for  selective  transmissions. 


SECTION  VI 

Radial  Roller  Bearings  with  Solid  Rollers    ♦ 

Solid  roller  journal  bearings  divide  into  two  classes,  the  first  using  cylindrical 
rollers,  and  the  second  conical  rollers.  Both  have  their  extensive  applications 
In  general,  the  cylindrical  roller  is  long  and  small  in  diameter  compared  with 
its  length,  while  the  conical  roller  is  usually  short  and  with  a  lesser  difference 
between  its  length  and  diameter.  The  cylindrical  roller  is  used  on  machinery 
in  general,  while  the  conical  roller  has  had  its  special  extensive  application  in 
automobile  practice  A  newer  development  in  the  use  of  the  cylindrical  roller 
is  for  the  bearings  of  street  railway  car  axles. 

Figs.    ii8    and  119  show   typical   designs   for   cylindrical   roller  journal 
bearings,  as  made  by  the  Standard  Roller  Bearing  Company.     The  first  is  a 


FIG.    118.      ROLLER  JOURNAL   PEDESTAL   BOX. 


Oil  Drain. 

Note:-  Space  Beariiigs  as  far  apart  as  possible 
aud  adjust  Collars  to  allow  desired 
Amount  of  End  Play  of  Shaft. 

FIG.    119.      ROLLER  JOURNAL  BEARING 
FOR  ELECTRIC   MOTOR. 


design  for  a  pedestal  box.  It  will  be  noted  that  the  rollers  are  comparatively 
small  in  diameter,  are  long,  have  a  ball  at  each  end  to  reduce  the  friction  from 
thrust  pressure,  are  held  together  in  a  cage  and  the  bearing  shell  is  fitted  into 
the  housing  by  a  spherical  seat  to  permit  of  self  alinement.  Another  feature 
worth  commenting  upon  is  the  felt  rings  at  the  end  to  exclude  dirt  and  keep 
in  lubricant. 

Fig.  119  shows  a  motor  bearing  having  two  rows  of  rollers.  This  bearing 
is  distinguished  from  the  one  that  precedes  it  by  the  fact  that  the  rollers  are 
shorter,  larger  in  diameter,  and  that  the  ends  have  angular  faces  to  take  up  a 
certain  amount  of  end  thrust.     These  oblique  faces  fit  between  corresponding 

217 


2l8 


BEARINGS  AND  THEIR  LUBRICATION 


shoulders,  so  that  the  motor  or  armature  shaft  can  float  endwise  a  short  distance 
and  be  checked  by  the  meeting  of  these  inclined  surfaces.  It  is  evident  that 
no  great  amount  of  end  thrust  can  be  successfully  carried.  The  ends  of  the 
faces  are  provided  with  a  felt  washer  and  grooves  to  check  the  outward  flow  of 
oil  and  the  entrance  of  dust. 


FIG.  1 20.   ROLLER  BEARING  FOR  ELECTRIC  CAR  TRUCK  AXLE. 

Fig.  120  shows  a  roller  bearing  designed  for  the  axles  of  electric  cars,  this 
particular  one  being  for  a  Halsey  truck.  The  journal  is  4  1/4  in.  in  diameter 
and  9  in.  long.  It  is  covered  by  a  steel  sleeve,  upon  which  the  rollers  bear. 
The  details  of  design,  including  provision  to  exclude  dust  and  provide  means 
for  introducing  lubricant  and  storage  place  for  the  lubricant  are  plainly  shown. 


FIG.  121.   ANOTHER  FORM  OF  ROLLER  BEARING  FOR  ELECTRIC  CAR  TRUCK  AXLE. 

At  the  end,  in  a  vertical  position,  is  a  small  plain  roller  thrust  bearing  to  take 
the  end  thrust  of  the  axle. 

Another  design  of  a  roller  bearing  for  this  purpose  is  shown  by  Fig.  121,  in 
which  there  are  two  sets  of  rollers  instead  of  one,  otherwise  the  general  details 
of  the  bearing  design  are  not  different,  except  that  this  has  a  large  pocket  at 
the  center  for  oil,  with  a  draining  plug  at  the  bottom.     This  design  makes 


RADIAL  ROLLER  BEARINGS  WITH  SOLID  ROLLERS 


219 


the  roller  bearing  interchangeable  with  plain  brass  lined  bearings  and  applicable 
to  the  standard  size  of  thrusts  in  which  the  latter  are  customarily  used. 

Bearings  of  this  kind  have  been  used  experimentally  to  quite  an  extent  in 
this  country  and  as  the  experimental  stage  can  be  assumed  to  have  been  passed, 
their  adoption  by  railways  may  be  looked  for  as  a  development  of  the  next 
few  years. 

FORMULA  FOR  CAPACITY  FOR  CYLINDRICAL  JOURNAL  ROLLER 

BEARINGS 

The  formula  for  capacity  for  solid  roller  journal  bearings  used  by  the 
Standard  Roller  Bearing  Company  is  as  follows: 

d'^nl 
130,000         =  capacity  in  pounds. 

Here  d  =  the  diameter  of  the  rollers  in  inches, 
n  =  the  number  of  rollers, 
Z  =  the  length  of  each  roller  in  inches, 
s  =  the  circumferential  speed  of  each  roller  in  feet  per  minute. 


FIG.  122.      CONICAL  ROLLER  JOURNAL  BEARINGS 
APPLIED  TO  AUTOMOBILE  HUBS. 


FIG.  123.  CONICAL  ROLLER  BEAR- 
ING AS  DIFFERENTIAL  BEARING  ON 
AN  AUTOMOBILE    REAR  AXLE. 


CONICAL  ROLLER  JOURNAL  BEARINGS 

Conical  roller  journal  bearings  have  had  their  most  extensive  use  in  auto- 
mobile practice.  Figs.  122,  123  and  124  are  from  designs  of  the  Standard 
Roller  Bearing  Company.  Fig.  122  showing  two  constructions  for  automo- 
bile hubs  having  two  single  rows  of  rollers.  Fig.  123  shows  a  bearing  of  this 
type,  used  as  a  differential  bearing  on  a  rear  axle.  Fig.  124  shows  in  section 
a  double  taper  roller  bearing. 


220 


BEARINGS  AND  THEIR  LUBRICATION 


FORMULA  FOR  CAPACITIES  FOR  TAPERED  OR  CONICAL  JOURNAL  BEARINGS 

The  load  carrying  capacity  in  pounds  is  given  by  the  following  formula, 
which  is  from  the  practice  of  the  Standard  Roller  Bearing  Company: 

130,000 =  capacity  in  pounds. 

In  which  d  =  the  mid-diameter  of  the  rollers  in  inches, 
n  =  the  number  of  rollers, 

I  =  the  contact  length  of  a  roller  with  its  bearing  washer  in  inches, 
5  =  the  mean  circumferential  speed  of  each  roller  in  feet  per  minute. 


FIG.    124.      DOUBLE    CONICAL   JOURNAL   BEARING. 


SECTION  VII 

Roller  Thrust  Bearings 

The  Standard  Roller  Bearing  Company  is  the  builder  of  a  patented  type  of 
plain  roller  thrust  bearings  that  has  been  adapted  to  a  wide  field  including 
situations  where  very  heavy  loads  are  to  be  sustained.  It  has  had  an  extensive 
application  for  water  turbine  generators;  in  fact,  its  first  use  for  heavy  loads  was 
in  connection  with  the  turbines  at  Niagara. 

The  design  consists  essentially  of  two  flat  plates  or  washers,  between  which 
is  a  bronze  cage  containing  a  number  of  cylindrical  rollers.     These  rollers  are 


PLAIN   ROLLER   STEP   BEARING   APPLIED   TO  A   CHEMICAL   FURNACE. 


comparatively  short  for  their  diameters,  have  spherical  ends  and  are  fitted  into 
radial  slots  in  the  cage.  The  outermost  roller  of  each  row  is  recessed  to  take 
a  ball,  which  is  held  in  a  hole  in  a  steel  band  surrounding  the  cage.  A  second 
band  then  surrounds  the  first  and  keeps  the  balls  in  position.  These  balls  are 
used  to  reduce  the  friction  due  to  end  thrust  from  centrifugal  force.  The 
washers  and  cages  are  made  either  split  or  solid  according  to  conditions. 


221 


222 


BEARINGS  AND  THEIR  LUBRICATION 


It  is  evident  that  the  motion  of  the  roller  is  a  combination  of  rolling  and 
sliding,  for  cylindrical  rollers  are  traveling  in  a  circular  pitch. 

These  bearings  are  made  for  all  kinds  of  load  and  service,  from  a  small  bear- 
ing adapted  to  a  7/8  in.  shaft  up  to  a  single  bearing  supporting  a  load  of  over 
2,000,000  lb. 

Figs.  125,  126  and  127  show  three  of  these  bearings.  The  first  is  for  a 
6  in.  shaft  and  was  designed  for  a  revolving  chemical  furnace.  Figs.  126  and 
127  are  of  larger  bearings,  the  latter  being  for  the  turbo-generators  at  McCall's 


FIG.  126. 


PLAIN    ROLLER   THRUST    BEARING    FOR   NINE-INCH   SHAFT, 
THROUGH   BEARING   PROPER. 


WITH    SHAFT   PASSING 


Ferry  on  the  Susquehanna  River.  Five  bearings  of  this  design  have  been 
installed  there  on  the  main  turbine  generators,  three  sustaining  a  load  each  of 
330,000  lb.,  and  two  a  load  each  of  410,000  lb.  The  size  of  the  bearing  can  be 
realized  when  the  shaft  diameter  is  seen  to  be  21  in. 

The  thrust  is  transmitted  to  the  bearing  through  a  collar  45  1/2  in.  in  di- 
ameter forged  integral  with  the  shaft. 

In  every  bearing  two  kinds  of  rollers  are  used — short  and  long.  These  are 
so  arranged  that  they  do  not  track,  but  cover  the  working  surfaces  of  the  washers 
so  that  the  latter  are  stressed  uniformly.     To  insure  perfect  alinement  of  the 


ROLLER  THRUST  BEARINGS 


223 


two  washers  it  is  customary  to  make  the  bottom  of  the  lower  one  hemispherical 
in  shape  and  seated  correspondingly  or  to  use  a  pair  of  leveling  plates. 

Lubrication  is  an  important  point.     In  the  larger  bearings  flooded  lubrica- 
tion is  maintained.     The  smaller  sixes  run  in  a  bath  of  oil. . 


Drain  Piping  Arrangemeut. 
FIG.    127.      PLAIN    ROLLER  THRUST  BEARING — TYPE   INSTALLED  AT   MCCALL'S   FERRY. 


FORMULA  FOR  CAPACITY 

The  formula  for  capacity  of  these  bearings,  based  finally  on  a  careful  analysis 
of  a  large  number  of  bearings  in  successful  use,  is  dependent  upon  the  roller 
diameter,  the  mean  circumferential  speed  of  the  roller  and  the  sum  of  their 
lengths.     This  formula  is: 

100,000— —-  =  load  capacity  in  pounds. 


2  24  BEARINGS  AND  THEIR  LUBRICATION 

n  which  D  =  the  diameter  of  the  rollers  in  inches; 

M  =  the  sum  of  the  lengths  of  all  the  rollers  in  inches; 

5  =  the  mean  circumferential  speed  of  the  rollers  in  feet  per  minute. 

The  value  100,000  of  the  constant  factor  is  a  minimum  and  thereby  gives  a 
minimum  value  for  the  capacity.  Owing  to  advantageous  conditions,  this 
numerical  factor  may  be  considerably  larger  in  some  designs. 

Experience  has  shown  that  good  practice  sets  a  maximum  limit  for  roller 
speeds  of  about  3000  r.  p.  m.  Similarly  the  maximum  circumferential  speed 
of  the  roller  is  fixed  at  about  3000  ft.  per  minute.  The  factors  of  a  uniform 
or  variable  speed  and  a  uniform  or  variable  load,  enter  to  modify  the  given 
numerical  constant. 

TEST  OF  BEARINGS  AT  McCALL'S  FERRY 

An  extensive  series  of  tests  has  been  run  on  the  bearings  at  McCall's  Ferry 
to  determine  the  friction  loss.     Valuable  data  have  been  obtained. 

A  word  of  explanation  in  regard  to  this  installation  is  necessary.  At  the 
time  of  testing  five  units  were  in  operation,  three  of  7500  and  two  of  10,000 
kilovolt  amperes  rating.  Each  turbine  has  a  rating  of  13,500  h.  p.,  with 
a  head  of  53  ft.  and  consists  of  two  Francis  wheels,  located  in  tandem  on  the 
same  shaft,  the  lower  one  being  below  the  tail  race  level.  The  generators  are 
three-phase,  25-cycle,  11,000  volts  at  a  speed  of  94  r.  p.  m.  The  first  unit 
at  the  time  of  this  writing  has  been  in  operation  for  over  one  year. 

Each  bearing  is  lubricated  by  flood  lubrication,  circulation  being  maintained 
by  gravity  with  the  supply  taken  from  an  elevated  tank.  The  overflow  from 
each  bearing  is  to  a  filter,  from  whence  it  is  pumped  into  the  storage  tank.  The 
filters  are  provided  with  cooling  coils  for  circulating  water,  but  these  are  seldom 
used  except  in  the  very  hottest  summer  weather.  Oil  enters  the  bearing 
through  the  lower  leveling  washer  into  the  space  outside  of  the  vertical  oil  guard. 
Thence  it  passes  upward  and  by  means  of  centrifugal  action  of  the  rollers  passes 
outward  through  the  openings  in  the  cage  and  between  the  rollers  themselves, 
and  overflows  at  a  level  about  midway  of  the  height  of  the  upper  bearing 
washer. 

In  case  the  oil  supply  should  be  cut  off  for  any  reason  whatever,  circulation 
will  still  be  automatically  maintained  by  the  pumping  action  of  the  rollers  and 
through  oil  grooves  in  the  lower  leveling  washer. 

Each  turbo-generator  has  three  babbitted  guide  bearings  and  two  lignum 
vitae  guide  bearings,  in  addition  to  the  roller  thrust. 

The  method  of  test  was  to  measure  accurately  the  rise  in  temperature  of  the 
oil  as  it  passed  out  of  the  bearing,  to  determine  carefully  the  amount  of  oil  in  the 
bearing,  to  find  the  unit  heat  capacity  of  the  oil,  and  to  determine  the  heat  loss 


ROLLER  THRUST  BEARINGS 


225 


226 


BEARINGS  AND  THEIR  LUBRICATION 


from  the  bearing  per  unit  of  time,  due  to  radiation,  dissipation  and  convection 
of  heat.  A  proper  handling  of  these  data  gave  the  amount  of  heat  liberated  in 
each  bearing  during  each  test.  Thus  it  involved  the  friction  of  the  rollers  rolling 
and  sliding  on  their  bearing  plates,  the  friction  of  the  rollers  between  themselves 
and  between  the  band;  in  fact,  all  of  the  frictional  losses  in  the  thrust  bearing 
itself.  Any  losses  taking  place  in  any  other  bearing  in  the  machine  in  no  wise 
affected  the  observations  or  results. 

Fig.  128  shows  the  graphic  log  of  a  test  of  one  of  the  more  heavily  loaded 
bearings.  It  will  be  seen  that  the  speed  was  varied  in  six  steps,  that  the  tempera- 
ture of  the  entering  oil  and  its  quantity  were  kept  practically  uniform.  The 
temperature  of  the  issuing  oil  is  seen  to  rise  with  the  various  steps  of  the  increas- 


S.9 
I 


r/ 

/ 

A 

/ 

/ 

/ 

/ 

/ 

ener 
Odd 

itorC 

"apac 

nrinn 

ity  =  l 
=4.in 

0,000 
100  Ih 

K.V.A. 

/ 

Normal  Speed  =  94  f\ 
Friction  Loss  in  Perce 

"ev.perMin. 
ntageof 

^afed  Generator  Out}. 

7ut=C 

.09% 

FIG.    129. 


0      10      20     .30     40     50     60     70     80     90      100 
Revolutions  per  Minute 

HORSEPOWER  LOSS  AT  VARIOUS  SPEEDS  IN  A  LARGE  PLAIN  ROLLER  BEARING. 
SEE    FIG.    128. 


ing  speed.  At  the  extreme  right  is  the  temperature  drop  per  unit  of  time  after 
the  turbine  was  shut  down  At  the  left  hand  end  are  data  giving  the  specific 
gravity  and  specific  heat  of  the  lubricating  oil. 

Fig.  129  shows  the  horsepower  loss  curve,  corresponding  with  the  test, 
graphically  shown  in  Fig.  128.  It  will  be  noticed  that  at  normal  speed  the  total 
loss  is  slightly  greater  than  12  h.  p.  Similarly  Fig.  130  shows  the  power 
loss  curve  for  the  same  kind  and  size  of  bearing,  but  with  a  load  of  330,000  lb., 
or  80,000  lb.  less  than  for  Fig.  129.  Here  will  be  noticed  that  the  horsepower 
loss  at  normal  speed  is  slightly  greater  than  lo  h.  p. 

Fig.  131  similarly  shows  a  horsepower  loss  curve  for  a  much  smaller  bearing, 
a  bearing  used  on  an  exciter.     Here  the  load  is  22,300  pounds. 

A  consideration  of  these  curves  shows  that  in  every  case  the  curve  is  a 
straight  line,  indicating  that  the  loss  is  directly  proportional  to  the  speed. 
Further  a  comparison  of  Figs.  129  and  130  shows  that  the  loss  is  roughly  pro- 
portional to  the  load. 


ROLLER  THRUST  BEARINGS 


227 


The  loss  for  the  two  larger  bearings  is  approximately  o.  i  per  cent,  of  the 
rated  generator  output. 

It  is  only  fair  to  state  that  the  bearing  tested  for  the  exciter  set  was  originally 


y 

' 

/ 

/ 

/ 

/ 

/ 

/ 

' 

/ 

/ 

/ 

c 

7ener 

afor 

Capacity- 

7500K.V.A. 

/ 

■Normal  Speed  =^94  Rev.perMin. 

/ 

7fGe 

aerator  Output 

=O.IC 

z 

/ 

0      10 


20     30    40     50     60     70     80 
Revolutions  per  Minute 


90     100 


FIG.    130. 


HORSEPOWER  LOSS  AT  VARIOUS  SPEEDS  IN  A  LARGE  PLAIN  ROLLER  BEARING 
SUPPORTING   A   LOAD    OF   330,000   POUNDS. 


designed  for  a  load  of  40,000  lb.,  and  thus  when  tested  was  working  at  only 
56  per  cent,  load,  therefore,  under  somewhat  disadvantageous  conditions. 
A  calculation  has  been  made  to  determine  the  coefficient  of  friction  for  the 


3 

X 

7 — 

/ 

\' 

y 

V 

7 

/ 

g- 

/ 

/ 

1' 

/ 

Lo 

nera 
ad  0! 

tore 

1  Ron 

opac 

rinn 

WOK 
nnih 

Y. 

/ 

Nc 
Fr 

)rmal  5peed=240  Rev.perMin. 
iction  Loss  in  Percentaqe^ 

/ 

of  Generator  Output  =-0.B3  Z 
111111 

FIG.    I3T. 


0      25      50     75     100     125     150     175    200    225   250 
Revolutions  per  Minute 

HORSEPOWER  LOSS   FOR  VARIOUS   SPEEDS   IN  A  PLAIN  ROLLER  BEARING 
APPLIED   TO  AN  EXCITER  SET. 


two  larger  bearings.  The  method  used  was  as  follows:  The  total  load  was 
divided  by  the  number  of  rollers  proportional  to  their  length.  This  gave  a  unit 
load  for  each  roller  in  the  bearing.     The  distance  from  the  mid-point  of  the 


228 


BEARINGS  AND  THEIR  LUBRICATION 


length  of  each  roller  to  th«  center  of  the  shaft  was  determined,  and  from  this 
was  calculated  the  moment  for  its  load.  These  moments  were  then  summed  for 
the  entire  bearing,  multiplied  by  the  average  angular  velocity  and  equated  to 
the  known  friction  loss  divided  by/  the  coefficient  of  friction. 

Solving  this  equation  gave  a  value  for  the  coefficient  of  combined  rolling 
and  sliding  friction  of  0.0012. 

CONICAL  ROLLER  THRUST  BEARINGS 

Table  47  is  taken  from  the  American  Machinist  for  February  13,  1908,  page 
239,  and  gives  the  proportions  for  conical  roller  thrust  bearings  as  made  by 
the  Standard  Machinery  Company.  The  design  of  these  bearings  consists  in 
making  the  roller  cone  apex  at  the  center  on  the  shaft  axis  and  the  angle  of  the 
cone  does  not  exceed  6  to  7  degrees.  If  the  angles  are  larger  than  these,  the 
outward  thrust  in  a  radial  direction  is  excessive  and  hence  causes  destruction 
of  the  outer  ends  of  the  rollers. 

Two  collars  form  the  rolling  surfaces  for  the  bearing,  one  being  stationary 
and  the  other  fastened  to  the  shaft  and  revolving  with  it.  The  inner  surfaces  of 
the  collars  are  made  conical  to  correspond  with  the  angle  of  the  rollers.  A  ring 
fits  tightly  over  the  bearing  and  serves  to  hold  the  rollers  in  position.  The 
collars,  as  well  as  the  ring,  are  of  high  carbon  tool  steel,  hardened  and  ground. 
The  rollers  are  of  medium  carbon  steel  and  spring  tempered. 


Diameter  of 

Number 

Area  of  bear 

Safe  pressure  on 

Safe  pressure  on 

shaft 

of  rolls 

ing  plate 

bear.ng,  speed  50 
rev. 

bearing,  speed  100 
rev. 

2-1/16  to  2-1/4 

30 

10.137 

19,000 

9,500 

3-1/ 1 6  to  3-1/4 

30 

20.862 

40,000 

20,000 

4-1/16  to  4-1/4 

30 

35 -^^97 

70,000 

35,000 

5-1/16  to  5-1/4 

30 

54-2 

108  000 

56,000 

6-1/16  to  6-1/2 

30 

78.017 

125  000 

62  000 

8-1/16  to  8-1/2 

32 

132.06 

200,000 

100  000 

9-T/16  to  g-1/2 

32 

162.98 

3^0,000 

150,000 

Table  47. — ^Proportion  or  Mossberg  Conical  Thrust  Bearings. 


INDEX 


Accuracy  of  ball  bearings,  189 
Acids  in  oils,  115 
Adhesion  of  Swedish  gages,  5 
Air  compressor  practice  in  bearing  pressures, 
66 
lubrication,  138 
Ajax  metal,  49 

Alining  devices  for  bearings,  90 
Alkalies  in  oils,  115 
Allan  red  metal,  47 
Allowable  bearing  pressures,  65 
Albys,  composition  of  for  bearings,  49,  50,  51 
of  copper,  tin  and  zinc,  51 
of  lead,  tin  and  antimony,  50 
for  metallic  packing,  61 
miscellaneous,  51 
physical  properties  of,  53 
soft,  for  bearings,  45 
of  tin,  copper  and  antimony,  50 
Analyses  of  bearing  metals,  49,  50,  51,  54 
Anchoring  of  babbitt  lining,  88,  147 
Animal  oils,  11 1 
Anti-attrition  metal,  56 

friction  metal,  48,  56 
Antimonial  lead,  48,  58 
Arc  of  contact,  effect  of  reducing,  134 
Atomizing  of  oil,  how  to  prevent,  169 
Automobile  ball  bearings,  201,  202 
bearing  metals,  60 
engine  bearing  pressures,  75 
engine  bearing  proportions,  87 
engine  bearing  rubbing  speeds,  80 
roller  bearings,  216,  219 

Babbitting,  invention  of,  139 

kinks,  167 
Babbitts,  45,  48,  49,  53,  54,  55 
Babbitt,  anchoring  of,  88,  147 

Genuine,  47 

German,  53,  54,  58 

Souther,  53,  54 

standard,  53,  54,  58 


Ball-and-socket  bearing,  typical  design  of,  155 
Ball  bearings,  accuracy  of,  189 

for  automobiles,  201,  202 

cages  for,  181,  188 

classification  of,  i 

cone  type,  188 

construction  of,  180—194 

double-row  radial,  186 

for  electric  motors,  199 

formulas  for,  177 

general  applications,  201 

limits  of  accuracy  for,  194 

lubrication  of,  204—207 

for  machine  tools,  195 

materials  for,  191 

mountings  for,  195—203 

proportions  of,  191 

radial,  capacity  of,  177 

thrust,  capacity  of,  178 

two-row  radial,  185,  187 
Ball  race  grooves,  curvature  of,  176 
Balls,  formula  for  unit  load  on,  174 
Bath  lubrication,  coefficients  of  friction  for, 

33,  34,  35,  36 
Bearing  inventions,  139 

metals,  See  Metals. 

pressures.  See  Pressures. 
Bearings  with  rolling  contact,  classification 
of,  2 

sliding  contact,  classification  of,  2 
Body  of  oils,  112 
Bores  of  bearings,  permissible  variations  in, 

90 
Brasses,  car,  wear  of,  30,  55 
Breaking-down  point  of  oil  film,  25 
Bronzes,  Allan,  48 

composition  of,  48,  49,  50,  50,  51,  54 

physical  properties  of,  53 

wear  of,  54,  55,  56 


Cages  far  ball  bearings,  181,  li 
Camelia  metal,  48 


229 


230 


INDEX 


Capacity  of  conical  roller,  journal  bearings, 
220 
conical  roller  thrust  bearings,  228 
cylindrical  roller,  journal  bearings,  219 
flexible  roller  bearings,  212 
plain  roller  thrust  bearings,  223 
radial  ball  bearings,  177 
rings  to  deliver  oil,  97 
roller,  journal  bearings,  219 
thrust  ball  bearings,  178 
Carbon  bronze,  48 
Car  box  metal,  49 

ring-oiling  design,  163 
brass  lining,  48 
brasses,  wear  of,  55 
truck  bearings,  158 
roller  bearings,  218 
Care  of  automobile  ball  bearings,  207 

bearings,  166 
Cast  iron  as  bearing  metal,  43 
Castor  oil,  in 
Centrifugal  oilers,  98 

Characteristics  of  a  good  bearing  metal,  41 
Classification  of  bearings,  i 
lubricating  devices,  93 
lubricants,  in 
Clearances  in  journal  bearings,  90 
Coefficient  of  friction,  definition  of,  6 
lateral,  39 

Morin,  10,  n,  12,  13,  14 
Stribeck,  175 
Coefficients  of  friction  of  ball  bearings,  175 
collar  bearings,  37 
flexible  roller  bearings,  212 

journal  bearings   minimum,  31 
leather  on  cast  iron,  14 
locomotive  values,  40 
metal  on  metal,  14 

oils  and  greases,  comparative,  113,  114 
plain  roller  thrust  bearings,  228 
radial  ball  bearings,  183 
roller  bearings,  175 
sliding  surfaces,  40 
step  bearings,  38 
thrust  ball  bearings,  189 
Tov^er  for  journal  bearings.    33,  34, 

35,  36 
unlubricated  surfaces,  Rennie,  10 
various  substances,  10 


Coefficients  of  wood  and  metal,  12,  13 

wood  on  wood,  11 
Collar  bearings,  coefficients  of  friction  of,  37 
moment  of  friction  of,  8 
work  of  friction  of,  8 
Combination    of    metals    for    journals    and 

bearings,  42 
Comparison  of  lubricating  methods,  20,  94 
Cone-type  ball  bearings,  188 
Conical  roller  journal  bearings,  219 

thrust  bearings,  228 
Construction  of  ball  bearings,  180-195 

journal  bearings,  141— 165 
Cooling  bearings  by  water  jacket,  no 
Cornish  bronze,  48 
Curvature  of  ball  race  grooves,  176 
Cylinder  lubrication,  quantity  of  oil  for,  116 
oil,  specifications  for,  130,  131 
(dark)  specification  for,  125 
(filtered)  specification  for,  126 
Cyprus  bronze,  53,  54,  56 

Damascus  bronze,  49 

DeLaval  steam  turbine  bearings,  156 

Delta  metal,  48 

Demo  bronze,  53,  54 

Design  of  ball  bearings,  174 

journal  bearings,  86—110 

journal  bearings  cooled  by  forced  lubrica- 
tion 109 
without  artificial  cooling,  108 
with  water  jacket  for  cooling,  no 

knife-edge  bearings,  136 

oil  grooves,  loi 

oil  rings,  98 

sliding  surfaces,  132 
Designs,  typical  for  journal  bearings,  141— 166 
Devices  for  anchoring  babbitt,  89,  147 

lubricating,  93 
Diameter  and  length  of  bearings,  ratio  of,  86 
Distribution  of  temperature  in  a  bearing,  99 
Dissipation  of  heat,  102 
Double-row  radial  ball  bearing,  186 
Driller  ball  bearing  mountings,  197 
Dry  surfaces,  laws  of  friction  for,  15 

Ex  B.  metal,  49,  58 

Electrical    machinery    practice    in    bearing 
pressures,  68 


INDEX 


231 


Electric  generator  bearings,  141 
Electric  motor  ball  bearing  mounting,  197 
journal  bearings,  141 
roller  bearings,  219 
".  ngi  e  oil,  specifications  of,  124,   127,  128, 

130 
E:  gines,  reciprocating,  type  plan  for  forced 
lubrication,  154 
steam,  lubrication  of  cylinders  of,  116 

Fiber  for  bearings,  64 
Filling  metal,  56 

Film  of  oil,  breaking  down  point  of,  25 
pressure  of,  25 
suction  of,  27 
thickness  of,  24 
Filtration  of  oil,  120 
Fits  of  bearings,  91,  92 
Flat  wearing  surfaces,  133 
Flexible  roller  bearings,  209 
Floating  sleeve  steam  turbine  bearing,  152 
Flooded  lubrication  for  machine  tools,  146 
quantity  of  oil  used,  115 
steam  turbine  bearing  arranged  for,  153 
Forced  lubrication,  94 
chart  of,  107 

design  of  bearings  for,  109 
point  to  introduce  oil,  94 
quantity  of  oil  used,  115 
type  plan  for  reciprocating   engines, 

154 
vertical    steam    turbine    bearing    ar- 
ranged for,  150 
Formulas  for  moment  of  friction,  8 

work  of  friction,  8 
Friction,    coefficients    for    journal    bearings, 

33.  34,  35,  36 
of  collar  bearings,  37 
of  step  bearings,  38 
coefficient  of,  lateral,  39 
definition  of  coefficient  of,  6 
formulas  for  moment  of,  8 

work  of,  8 
laws  of,  for  dry  surfaces,  15 

for  well  lubricated  surfaces,  19 
lowest  coefficients  for  journal  bearings, 

31 
of  lubricated  surfaces,  15 
of  poorly  lubricated  surfaces,  16 


Friction  of  radial  ball  bearings,  183 

of  rest,  6 

rolling,  173 

of  sliding  surfaces,  40 

of  thrust  ball  bearings,  189 

of  unlubricated  surfaces,  9 

of  well  lubricated  surfaces,  17 
Fractional  loss  in  bearings,  formula  for,  106 

resistance  and  load,  relationship  of,  21 

Gas  engine  oil,  specification  for,  126,  127 
practice  in  bearing  pressures,  69—73 
Generator,  electric,  ball  bearings,  199 

journal  bearings,  141 
Genuine  babbitt,  47 
German  babbitt,  53,  54,  58 

practice  in  bearing  pressures,  82 
railroad  practice  in  use  of  bearing  metals, 
58 
Graney  metal,  49 
Graphite  bearing  metal,  48 
as  lubricant,  iii,  112 
in  bearings,  64 
Grease  and  oil,  comparison  of  friction  of,  113, 

114 
Greases,  tests  for,  121 

Grinder  spindle  ball  bearing  mounting,  195 
196 
bearings,  137,  142,  148 
Grooves  for  oil,  100 

in  ball  bearing  races,  curvature  of,  176 

Hard  lead,  49 
Harrington  bronze,  49 
Heat  liberated,  formulas  for,  8 
radiation  of,  102 
unit  factors  for,  103 
Horizontal  steam  turbine  bearings,  151,  156 
Hot  box,  how  to  cure,  168 
Hyatt  roller  bearings,  209 

Influence  of  bearing  metals  on  friction,  19 
Inventions,  three  important  ones,  139 

Journal  bearings,  classification  of,  2 
clearances  of,  90 

coefficients  of  friction  for,  33,  34, 35,  36 
design  of,  106 
heat  radiation  of,  103 
moment  of  friction  of,  8 


232 


INDEX 


Journal  bearings,  tests  of,  8i,  103, 104, 105, 107 
wear  of,  56 
work  of  friction  of,  8 
design  of,  86 

diameters,  permissible  variations  in,  90 
oil  grooves,  loi 

Kingsbury  thrust  bearing,  159 
Knife-edge  bearings,  136 

Lard  oil,  coefficients  of  friction  for,  34,  35, 
114 
specifications  for,  122 
test  for,  1 23 
Lateral  friction,  coefficient  of,  39 
Laws  of  dry  friction,  15 

friction,  Morin,  9 

for  well  lubricated  surfaces,  19 
Length  and  diameter  of  bearings,  ratio  of.  86 
Limits  for  bearings,  91,  92 

journals,  91,  92 

radial  ball  bearings,  194 
Line  shaft  bearing,  typical  design  of,  155 
Lining  metal,  anchoring  of,  89,  147 
Load  and  f rictional  resistancfe,  relationship  of, 
21 

carrying  capacity.  See  Capacity. 
Locomotive  eccentric  strap,  161 

rod  stubs,  161,  162 

slide  valves,  coefficient  of  friction  for,  40 

truck  bearings,  158 
Loss  of  power  in  bearings,  8 
Lubricated  surfaces,  friction  of,  15 
Lubricants  for  ball  bearings,  204—207 

classification  of,  in 
Lubricating  devices,  classification  of,  93 

greases,  classification  of,  in 

methods,  comparison  of,  20,  94 

oils,  classification  of,  in 
filtration  of,  1 20 

specifications  for,  123 
Lubrication,  air,  138 

forced,  chart  of,  107 

of  ball  bearings,  204—207 

of  flexible  roller  bearings,  213 

of  sliding  surf  aces,  133 

of  steam  engine  cylinders,  116 

of  steam  turbines,  120 

theory  of,  18 


Machine  oil,  specifications  for,  129 
tool  ball  bearings,  195 

bearings,  145 
tool,  temperature  of  bearings  of,  84 
Magnolia  metal,  48,  57 
Manganese  bronze,  49,  53 
Materials  of  ball  bearings,  191 
Mercury  bearings,  64 
Metal,  anchoring  of,  89,  147 

on  metal,  coefficients  of  friction  for,  14 
Metal  and  wood,  coefiicients  of  friction  for, 

12,  13 
Metallic  packing,  61 
Metals,  See  Babbitts  and  Bronzes. 
Metals  for  automobile  bearings,  60 

bearings,  analyses  of,  48,  49,  50,  51,  54 
characteristics  of,  41 
physical  properties  of,  53 
selection  of,  88 
white,  45 
combination  of,  for  bearings  and  jour- 
nal, 42 
composition  of,  48,  49,  50,  51,  54 
experiment  with,  for  bearings,  44 
influence  of,  on  friction,  19 
used  in  DeLaval  turbine,  58 
by  Pennsylvania  R.  R.  Co.,  58 
by  U.  S.  Navy,  56,  57 
by  U.  S.  War  Department,  56 
variations  in,  for  bearings,  63 
wear  of,  in  bearings,  54 
Mineral  grease,  coefficients  of  friction  for,  2>3 

oil,  coefficient  of  friction  for,  33,  114 
Moment  of  friction,  8 
Morin's  friction  experiments,  9 

laws  of  friction,  9 
Motion,  friction  of,  6 
Motor  ball  bearing  mountings,  197,  198,  199 

journal  bearings,  141 
Mountings  of  ball  bearings,  195—203 

Naval  practice  in  bearing  pressures,  66 

white  brass,  57 
Navy  bearing  metals,  56 

specifications  for  oils,  123 

Oil,  coefficients  of  friction  for  various  kinds, 

2>2>,  34,  35,  36 
delivery  capacity  of  rings,  97 


INDEX 


233 


Oil,  feed,  rates  of,  for  flooded  lubrication,  115 
forced  lubrication,  107,  115 
steam  engine  cylinders,  116 
steam  turbine  steps,  120 
film,  breaking  down  point,  25 
point  of  maximum  pressure,  27 
minimum  pressure,  27 
nearest  approach,  27 
pressure  of,  25 
suction  of,  27 
thickness  of,  film  of,  24 
filtration  of,  120 
and  grease,  comparison  of  friction  of ,  1 13, 

114 
grooves,  100 

how  to  prevent  atomizing  of,  169 
hole  with  left  washer,  95 
point  to  introduce  in  forced  lubrication 

94 
quantities  used,  107,  115,  116,  120 
ring  bearing,  temperature  rise  of,  104, 

105 
rings,  design  of,  97 
Oils,  properties  of,  112 

specifications  for,  122 
Olive  oil,  coefficient  of,  friction  for,  35,  114 


Practice  in  use  of  bearing  metals,  45,  56 
Pressure  and  velocity,  product  of,  85 

maximum  point  in  film,  27 

minimum  point  in  film,  27 

in  oil  film,  25 
Pressures  in  automobile  engine  bearings,  75— 

79 
bearings,  air  compressor  practice,  66 
electrical  machinery  practice,  68 
gas  engine  practice,  69,  74 
German  practice,  82 
knife-edge  bearings,  137 
maximum  safe  for  perfect  film  lubri- 
cation, 81 
naval  practice,  66 
railroad  practice,  65 
rolling  mill  practice,  68 
steam  engine  practice,  66 
flexible  roller  bearings,  212 
Product  of  pressure  and  velocity,  85 
Proportions  of  automobile  engine  bearings 
87 
of  journal  bearings,  86 
radial  ball  bearings,   191 
Provisions  for  dissipating  heat,  102 
lubricating,  93 


Pad  lubrication,  coefficients  of  friction  for,  20, 

36,  94 
Parson's  white  brass,  54,  57 
Pedestal  bearings,  typical  design  of,  141 
Pennsylvania   Railroad   Company's   bearing 

metals,  58 
Permissible  rise  in  temperature,  83 
Phosphor  bronze,  49,  53,  54,  56,  58,  60 
Physical  properties  of  bearing  metals,  53 
Pivot  bearings,  moment  of  friction  of,  8 
work  of  friction  of,  8 
step  bearings,  coefficient3  of  friction  of, 
38 
Plain  roller  thrust  bearings,  221—228 

test  of,  224 
Plastic  bronze,  53,  54,  56 
Plumbic  bronze,  53,  54 
Point  to  introduce  oil  with  forced  lubrication, 
94 
of  nearest  approach  of  bearing  and  jour- 
nal, 27 
Power  lost  in  friction,  8 


Quantities  of  oil  used  for  lubricating,  115 

Race  grooves  in  ball  bearings,  curvature  of 

176 
Radial  ball'  bearings,  cages  for,  181 

capacity  of,  177 

doubje-row,  186 

friction  of,  183 

limits,  194 

proportion    of,  192 

tolerances,  194 

tests  of,  184 

two-row,  185,  187 

types  of,  180 
Radiation  of  heat  from  bearings,  102 
Railroad  practice  in  bearing  pressures,  65 

use  of  bearing  metals,  58,  59 
Rape  oil,  coefficients  of  friction  for,  36,  37 

114 
Ratio  of  bearing  length  to  diameter,  86 
Reciprocating  engines,  type  plan  for  forced 
lubrication  of,  154 


234 


INDEX 


Red  brass,  6i 

metal,  Allan,  48 
Reducing  arc  of  contact,  effect  of,  134 
Rest,  friction  of,  6,  16 
Ring  oiling,  95 

car  box,  163 
invention  of,  140 
Rise  in  temperature,  permissible,  83 
Roller  bearings,  classification  of,  2 
flexible  rollers,  209-216 
journal  bearings,  conical  rollers,  219 

cylindrical  rollers,  217 
plain  thrust  bearings,  221 
thrust  bearings  with  conical  rollers,  228 
Rollers,  unit  loads  on,  174 
Rolling  friction,  173 

mill,  practice  in  bearing  pressures,  68 
Rubbing  speed  in  automobile  engine  bear- 
ings, 80 
relation  between  and  rise  in  tempera 

ture,  104,  105,  107 
with  relation  to  pressure,  81 
velocity,  permissible,  82 
Rules  for  bearing  design,  106 
Running  fits  for  shafts,  91 

Salgee  metal,  48 
Scale  bearings,  136 
Selection  of  bearing  metal,  88 
Seller's  bearing,  invention  of,  140 
Shonberg  M.  M.  metal,  54,  57 
Slack  in  bearing,  how  to  take  up,  166 
Sleeve  bearings,  typical  design  of,  143 
Sliding  fits  for  shafts,  91 

friction,  theory  of,  5 

surfaces,  design  of,  132 
lubrication  of,  133 
Solid  lubricants,  iii 

roller  bearings,  217—228 
Souther  babbitt,  53,  54 
Specifications  for  lard  oil,  122 

oils,  U.  S.  Navy,  123 

U.  S.  War  Department,  124 
Sperm  oil,  coefficients  of  friction  for,  34    114 
Spindle  bearings  for  traverse  grinder,  137 
Standard  babbitt,  53,  54,  58 
Steam  engine  bearing  pressures,  66 
lubrication  of  cylinders  of,  116 
piston  bearing  metal,  47 


Steam  turbine  bearings,  150 

quantity  of  oil  used  in,  120 
Step  bearings,  coefficients  of  friction  for,  38 

for  vertical  shaft,  typical  design  of,  155 

quantity  of  oil  used  in,  120 
Suction  of  oil  film,  27 
Sulphur  as  lubricant,  in,  168 
Swedish  gages,  adhesion  of,  5 
Syphon    lubrication,    coefficients   of    friction 
for,  20,  37 

Talc  as  lubricant,  in,  168 
Tallow  oil,  in 

Temperature  distribution  in  a  bearing,  99 
rise  of  journal  bearings,  104,  105,  107 
permissible,  83 
Temperatures  in  machine  tool  bearings,  84 
Tender  boxes,  159 

ball  bearings,  183,  189 
flexible  roller  bearings,  212 
journal  bearings,  56,  81,  103,  104,  105, 
107 
Test  of  plain  roller  thrust  bearings,  224 
Testing  machine  bearings,  136 
Tests  for  greases,  121 

lard  oil,  123 
Textile  machinery  bearings,  164 
Theory  of  lubrication,  18 
Thickness  of  oil  film,  24 
Thrust  ball  bearings,  cages  for,  188 
capacity  of,  178 
friction  of,  189 
bearings,  conical  roller,  228 
Kingsbury,  159 
plain  roller,  221 
steam  turbine,  150 
Tobin  bronze,  49 
Tolerances  in  journal  bearings,  90 

radial  ball  bearings,  194 
Traverse  grinder  spindle  bearings,  137 
Turbines,  steam,  bearing  metals  of,  58 
bearings  of,  150 
quantity  of  oil  used  in,  1 20 
Two-row  radial  ball  bearings,  185-187 
Typical   constructions   of   journal   bearings 
141-165 
designs  of  journal  bearings,  141-165 

Unit  pressure,  formula  for,  7 


INDEX 


235 


Unlubricated  friction,  laws  of,  15 
surfaces,  friction  of,  9 

Valves,  coefficient  of  friction  for  locomotive, 

40 
Variations  in  bearing  metals,  6^ 
Vegetable  oilS;  1 1 1 
Velocity  and  pressure,  product  of,  35 
of  rubbing,  82 
in  automobile  engine  bearings,  80 

in  relation  to  coefficient  of  friction, 

22 
Vertical  steam  turbine  bearings,  150 
Viscosity  of  oils,  112 

War  Department  bearing  metals,  57 
specifications  for  oils,  1 24 


Water  as  lubricant,  112 

cooled  steam  turbine  bearing,  151 

jacket  for  cooling,   no 
Wear  of  bronzes,  54,  55,  56 

filling  metal,  56 
Wearing  surfaces,  fiat,  design  of,  133 
Weighing  machine  bearings,  136 
Whale  oil  as  lubricant,  in 
White  brass,  60 

metal,  48 

metals  for  bearings,  45 
Wood  for  bearings,  63 

and  metal,  coefficients  of  friction  for,  1 2, 

13 
on  wood,  coefficients  of  friction  for,  11 
Work  of  friction,  7,  8. 


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